UNIVERSITY    OF  CALIFORNIA 


ARCHITECTURAL 


GIFT  OF 
Mrs.  George  Beach. 


PREFACE 


PREFACE 


BY 

G.  F.  GEBHARDT 

PROFESSOR    OF    MECHANICAL    ENGINEERING,    ARMOUR 

INSTITUTE    OF    TECHNOLOGY 

CHICAGO,    ILL. 


THIRD  EDITION,  REVISED  AND  ENLARGED 
FIRST    THOUSAND 


NEW   YORK 

JOHN    WILEY   &   SONS 

LONDON:    CHAPMAN   &   HALL,   LIMITED 

1910 


:Jr| 


COPYBIGHT,  1908,  1910, 
BY 

G.   F.   GEBHARDT 


Stanhope  ipreaa 

t.   H.  GILSON     COMPANY 
BOSTON.      U.S.A. 


PREFACE 


THIS  book  is  the  outcome  of  a  series  of  lectures  delivered  to  the 
Senior  class  of  the  Armour  Institute  of  Technology,  Chicago,  111. 
It  is  primarily  intended  as  a  text-book  for  engineering  students, 
but,  it  is  hoped,  will  also  be  of  interest  to  practicing  engineers. 

The  field  embraced  by  the  title  is  a  large  one  and  it  has  been 
necessary  to  limit  the  treatment  to  essential  elements.  Much  of  the 
matter  contained  in  the  author's  original  notes,  including  that  relat- 
ing to  steam  engine  design,  valve  gears,  steam  boiler  design,  and  the 
like,  has  therefore  been  omitted.  The  numerous  references  appear- 
ing throughout  the  text  and  the  appended  bibliographies,  which 
have  been  carefully  compiled,  are  depended  upon  to  extend  the 
scope  of  the  work.  The  standard  codes  of  the  American  Society  of 
Mechanical  Engineers  for  conducting  engine  and  boiler  trials  are  in 
frequent  demand  by  engineers  and  have  therefore  been  included  as 
an  appendix. 

Authorities  have  been  freely  consulted  and  extensive  use  made  of 
current  engineering  literature,  due  acknowledgment  being  made  by 
footnote  or  reference  whenever  possible. 

The  matter  included  is  representative  of  American  practice  and  no 
effort  has  been  made  to  include  any  other  except  in  a  few  special 
cases. 

The  author  wishes  to  express  his  obligations  to  Prof.  Raymond 
Burnham  for  many  valuable  suggestions  and  corrections,  and  to 
Mrs.  Julia  Beveridge,  librarian  at  Armour  Institute,  for  assistance 
in  compiling  references. 


812370 


PREFACE   TO   SECOND   EDITION 


A  NUMBER  of  additional  changes  have  been  made  to  bring  this 
work  into  accord  with  more  recent  practice.  All  the  typographical 
and  other  errors  discovered  in  the  first  edition  have  been  corrected. 


PREFACE   TO   THIRD   EDITION 


ALL  obsolete  matter  has  been  discarded,  considerable  new  material 
has  been  added  throughout  the  book,  and  many  of  the  chapters  have 
been  entirely  rewritten. 


iv 


CONTENTS 


PAGE 

CHAPTER  I.  —  ELEMENTARY  STEAM  POWER  PLANTS.  ..-...' .. .  1-13 

1.  General 1 

2.  Elementary  Non-Condensing  Plant 2 

3.  Non-Condensing  Plant.     Exhaust  Steam  Heating ..'.-.  I  . 5 

4.  Elementary  Condensing  Plant 7 

5.  Condensing  Plant  with  Full  Complement  of  Heat-Saving  Appliances,. .,  10 

CHAPTER  II.  —  FUELS  AND  COMBUSTION ,  14-67 

6.  General.  .  .. 14 

7.  Classification  of  Fuels. 14 

8.  Solid  Fuels 14 

9.  Coal .......,,,.,....,... 15 

10.  Anthracite 15 

11.  Semi- Anthracite ................'.. 16 

12.  Semi-Bituminous ..........'.... 16 

13.  Bituminous 16 

14.  Lignite 17 

15.  Peat  or  Turf. ..........::::.:.: is 

16.  Wood,  Straw,  Sawdust,  Bagasse,  Tanbark 19 

17.  Composition  of  Coal 24 

18.  Combustion 24 

19.  Temperature  due  to  Combustion 27 

20.  Air  Required  for  Combustion 28 

21.  Calorific  Value  of  Coal 31 

22.  Heat  Losses  in  Burning  Coal 32 

23.  Loss  in  the  Dry  Chimney  Gases 33 

24.  Loss  due  to  Incomplete  Combustion 36 

25.  Loss  of  Fuel  through  Grate 37 

26.  Superheating  the  Moisture  in  the  Air 37 

27.  Loss  due  to  Moisture  in  the  Fuel 37 

28.  Loss  due  to  Presence  of  Hydrogen  in  the  Fuel 38 

29.  Loss  due  to  Visible  Smoke 38 

30.  Radiation  and  Minor  Losses 39 

31.  Size  of  Coal 39 

32.  Washed  Coal 40 

33.  Purchasing  Coal : 42 

34.  Powdered  Coal 44 

35.  Depreciation  of  Powdered-Coal  Furnaces 44 

36.  Storing  Powdered  Fuel *. 45 

37.  Rate  of  Combustion  with  Powdered  Coal 45 

38.  Cost  of  Pulverizing  Coal 45 

39.  Efficiency  of  Powdered-Coal  Furnaces 46 

v 


vi  CONTENTS 

CHAPTER    II  —  Continued  PAGE 

40.  Furnaces  for  Burning  Powdered  Coal 47 

41.  Draft  for  Powdered  Fuel 47 

42.  Types  of  Powdered-Coal  Burners 47 

43.  Pinther  Apparatus  for  Burning  Powdered  Coal 48 

44.  Schwartzkopff  Apparatus  for  Burning  Powdered  Coal 49 

45.  Aero-Pulverizer  Apparatus  for  Burning  Powdered  Coal 49 

46.  Triumph  Apparatus  for  Burning  Powdered  Coal 50 

47.  Fuel  Oil 51 

48.  Chemical  and  Physical  Properties  of  Fuel  Oil 52 

49.  Efficiency  of  Boilers  with  Fuel  Oil 54 

50.  Comparative  Evaporative  Economy  of  Oil  and  Coal 55 

51.  Types  of  Oil  Burners 55 

52.  Furnaces  for  Burning  Oil  Fuel 59 

53.  Air  vs.  Steam  as  an  Atomizing  Medium 60 

54.  Oil  Pressure , 63 

55.  Oil  Storage  and  Transportation 64 

56.  Conclusions  of  the  U.  S.  Naval  Liquid  Fuel  Board 65 

57.  Gaseous  Fuels 66 

CHAPTER   III.  —  BOILERS 66-123 

58.  General 68 

•59.    Classification 68 

60.  Vertical  Tubular  Boilers 68 

61.  Fire-Box  Boilers 71 

62.  Scotch-Marine  Boilers 72 

63.  Robb-Mumford  Boiler 73 

64.  Horizontal  Return  Tubular  Boiler 74 

65.  Babcock  &  Wilcox  Boiler 81 

66.  Heine  Boiler 82 

67.  Wickes  Vertical  Water-Tube  Boiler 82 

67a.    Parker  Boiler 86 

68.  Stirling  Boiler 87 

69.  Unit  of  Evaporation 88 

69a.    Heat  Transmission 90 

70.  Heating  Surface 92 

71.  Horse  Power  of  a  Boiler 93 

72.  Grate  Surface 95 

73.  Boiler  and  Furnace  Efficiency 98 

74.  Boiler  Performances 99 

75.  Effect  of  Capacity  on  Efficiency 104 

76.  Thickness  of  Fire 109 

77.  Influence  of  Initial  Temperature  on  Efficiency Ill 

78.  Cost  of  Boiler  and  Settings 112 

79.  Selection  of  Type 112 

80.  Grates 114 

81.  Rocking  Grates 115 

82.  Blow-Offs 116 

83.  Dampers 118 

84.  Water  Gauge 119 


CONTENTS  vii 

CHAPTER  III  —  Continued  PAGE 

85.  Fusible  Plugs 121 

86.  Mechanical  Tube  Cleaners 121 

CHAPTER  IV.  —  SMOKE  PREVENTION,  FURNACES,  STOKERS 124-151 

87.  General 124 

88.  Mechanical  Stokers 125 

89.  Chain  Grates 126 

90.  Step  Grates,  Front  Feed 130 

91.  Step  Grates,  Side  Feed , 136 

92.  Underfeed  Stokers 138 

93.  Down-Draft  Furnaces 139 

94.  Sprinkling  Furnaces 141 

95.  Dutch  Ovens 141 

96.  Twin-Fire  Furnace 142 

96a.   Chicago  Settings 143 

97.  Wooley  Smokeless  Furnace 149 

98.  Kent's  Wing-Wall  Furnace 149 

99.  Burke's  Smokeless  Furnace 151 

100.  Admission  of  Air  above  Fire 151 

101.  Cost  of  Stokers  and  Furnaces 151 

CHAPTER   V.  —  SUPERHEATED  STEAM;  SUPERHEATERS 152-180 

102.  General 152 

103.  Economy  of  Superheat 153 

104.  Limit  of  Superheat 154 

105.  Specific  Heat  of  Superheated  Steam 155 

106.  Types  of  Superheaters 162 

107.  Babcock  &  Wilcox  Superheater 163 

108.  Stirling  Superheater 164 

109.  Foster  Superheater 165 

110.  Independently  Fired  Superheaters 166 

111.  Materials  for  Superheaters 170 

112.  Extent  of  Superheating  Surface 170 

113.  Performance  of  Superheaters 174 

113a.   Properties  of  Superheated  Steam 180 

CHAPTER   VI.  —  COAL  AND  ASH-HANDLING  APPARATUS 181-206 

114.  General 181 

115.  Coal  Storage 181 

116.  Coal  Conveyors 183 

117.  Hand  Shoveling 183 

118.  Bucket  Conveyors 184 

119.  Belt  Conveyors 192 

120.  Elevating  Tower,  Hand-Car  Distribution 193 

121.  Overhead  Storage,  Bucket  Hoist 195 

122.  Elevating  Tower,  Cable-Car  Distribution 196 

123.  "  Vacuum  "  Ash-Handling  System 196 

124.  Cost  of  Handling  Coal  and  Ashes 201 

125.  Coal  Hoppers 202 

126.  Coal  Valves .  ,  .205 


vm  CONTENTS 

PAGE 
CHAPTER   VII.  —  CHIMNEYS 207-244 

127.  Chimney  Draft ,  *  • 207 

128.  Chimney-  Formulas .. . 212 

129.  Height  of  Chimneys  for  Boilers  Using  Oil  Fuel 218 

130.  'Classification  of  Chimneys 218 

131.  -Guyed-  Chimneys 219 

132.  'Self-Sustaining  Steel  Chimneys 219 

133.  Thickness  of  Plates • 220 

134.  Riveting 223 

135.  Stability  of  Steel  Chimneys 223 

136.  Brick  Chimneys 224 

137.  Thickness  of  Walls .  . 226 

138.  Core  and  Lining . 230 

139.  Materials  for  Brick  Chimneys 230 

140.  Stability  of  Brick  Chimneys . 231 

141.  'Custodis  Radial  Brick  Chimney.  . 234 

142.  Steel-Concrete  Chimneys : 234 

148.    Breeching.  . 240 

144.  Chimney  Foundations ...-.; 240 

145.  Chimney  Efficiencies 241 

146.  "  Cost  X)f  Chimneys 243 

CHAPTER   VIII.  —  MECHANICAL  DRAFT 245-266 

147.  General. 245 

148.  Steam  Jets . ... 245 

149.  Parsons  Smokeless  Furnace 248 

150.  Heinrich  Smokeless  Furnace 249 

151.  Fan  Draft 249 

152.  Theory  of  Fans 252 

153.  Determination  of  the  Size  of  Fan 258 

154.  Chimney  vs.  Mechanical  Draft 261 

155.  Balanced  Draft 264 

CHAPTER   IX.  —  STEAM  ENGINES 267-326 

156.  Introductory . .. 267 

157.  The  Ideal  Engine . 267 

158.  Thermal  Efficiency  of  the  Actual  Engine 273 

159.  Mechanical  Efficiency 275 

160;  Heat  Losses  in  the  Actual  Engine 278 

161.  Loss  due  to  Moisture  in  the  Steam  at  Admission 278 

162.  Loss  -due  to  Leakage 279 

163.  Loss  due  to  Cylinder  Condensation 279 

164.  Loss  due  to  Clearance  Volume 281 

165.  Loss  due  to  Incomplete  Expansion  and  Compression 282 

166.  Loss  due  to  Wire  Drawing 284 

167;  Loss  due  to  Friction 284 

168.  Effect  of  Increasing  Boiler  Pressure 286 

169.  Receiver-Reheaters 287 

170.  Jackets , . . 288 

171*    Single  and  Double  Acting  Engines 290 

172.  •  High-  and  Low  Speed  Engines 290 


CONTENTS  ix 

CHAPTER   IX  — Continued  PAGE 

173-4.   High-Speed  Single- Valye  Engines 291 

175.  High-Speed  Multi-Valve  Engines 297 

176.  Medium  and  Low  Speed  Engines 299 

177.  Compound  Engines...  .  . 300 

178.  Triple  and  Quadruple  Engines 305 

179.  Influence  of  Condensing 307 

180.  Throttling  vs.  Automatic  Cut-Off 310 

181.  Influence  of  Superheat 313 

182.  Binary  Vapor  Engines , 321 

183.  Cost  of  Engines , « 326 

CHAPTER    X.  —  STEAM  TURBINES 327-396 

184.  Classification 327 

184a.    General  Elementary  Theory " 328 

185.  De  Laval  Turbine 331 

186.  Elementary  Theory,  De  Laval  Turbine 333 

187.  Terry  Turbine 345 

188.  Kerr  Turbine 346 

189.  Curtis  Turbine 350 

190.  Elementary  Theory,  Curtis  Turbine 358 

191.  Hamilton-Holzworth  Turbine 362 

192.  Westinghouse-Parsons  Turbine 365 

192a.    Allis  Chalmers  Turbine !  .  .  / 372 

193.  Elementary  Theory,  Parsons  Turbine 373 

194.  Low  and  Mixed  Pressure  Turbines 376 

195.  Advantages  of  the  Steam  Turbine 382 

196.  Simplicity 382 

197.  Economy  of  Space  and  Foundation 382 

198.  Absence  of  Oil  in  Condensed  Steam 384 

199.  Regulation 384 

200.  Overload  Capacity 384 

201.  Efficiency  and  Economy 384 

202.  First  Cost 389 

203.  Cost  of  Operation 392 

204.  Influence  of  Superheat 393 

205.  Influence  of  High  Vacua 395 

CHAPTER    XL  —  CONDENSERS 397-469 

206.  General 397 

207.  Function  of  the  Condenser 398 

208.  Classification  of  Condensers 400 

209.  Common  Jet  Condenser 401 

210.  Condensing  Water,  Jet  Condensers 404 

211.  Effect  of  Aqueous  Vapor  upon  the  Degree  of  Vacuum 405 

212.  Injection  Orifice 407 

213.  Volume  of  Condenser  Chamber 408 

214.  Injection  and  Discharge  Pipes 408 

215.  Siphon  Condensers 408 

216.  Size  of  Siphon  Condensers 409 

217.  Ejector  Condensers 410 


x  CONTENTS 

CHAPTER   XI  — Continued  PAGB 

218.  Barometric  Condensers 411 

219.  Water-Cooled  Surface  Condensers 416 

220.  Cooling  Water,  Surface  Condensers 420 

221.  Extent  of  Water-Cooling  Surface 421 

222.  Dry- Air  Surface  Condensers 428 

223.  Quantity  of  Air  for  Cooling  (Dry- Air  Condenser) 429 

224.  Saturated- Air  Surface  Condensers 430 

225.  Evaporative  Surface  Condensers 433 

226.  Location  and  Arrangement  of  Condensers 433 

226a.    Independent  System 434 

227.  Central  Condensing  Systems 439 

228.  High- Vacuum  Systems 441 

229.  Power  Consumption  of  Condenser  Auxiliaries 447 

230.  Cost  of  Condensers 450 

231.  Most  Economical  Vacuum 451 

232.  Choice  of  Condensers 452 

233.  Water-Cooling  Systems 453 

233a.   Cooling  Pond 454 

233b.    Spray  Fountain 455 

234.  Cooling  Towers 456 

235.  Parallel  Comparison  of  Fan  and  Natural-Draft  Cooling  Towers 460 

236.  Cooling-Tower  Calculations 460 

236a.   Hygrometry 468 

237.  Tests  of  Cooling  Towers 468 

CHAPTER   XII.  —  FEED-WATER  PURIFIERS  AND  HEATERS 471-522 

238.  General 471 

239.  Chemical  Purification 476 

240.  Boiler  Compounds 476 

241.  Use  of  Kerosene  and  Petroleum  Oils  in  Boiler  Feed  Water 477 

242.  Use  of  Zinc  in  Boilers 478 

243.  Methods  of  Introducing  Compounds 478 

244.  Weight  of  Compound  Necessary 478 

245.  Mechanical  Purification 479 

246.  Thermal  Purification 479 

247.  Purifying  Plants 480 

248.  Economy  of  Preheating  Feed  Water 484 

249.  Classification  of  Feed- Water  Heaters 485 

250.  Open  Heaters 486 

251.  Open  Heaters  and  Purifiers 489 

252.  Temperatures  in  Open  Heaters 489 

253.  Pan  Surface  Required  in  Open  Heaters 491 

254.  Size  of  Shell,  Open  Heaters 491 

255.  Classification  of  Closed  Heaters 492 

256.  Closed  Heaters,  Water-Tube 493 

257.  Closed  Heaters,  Steam-Tube 494 

258.  Heating  Surface,  Closed  Heaters 496 

259.  Heat  Transmission,  Closed  Heaters 497 

260.  Open  vs.  Closed  Heaters 504 

261.  Through  Heaters 505 


CONTENTS  xi 

CHAPTER    XII  —  Continued  PAGE 

262.  Induced  Heaters 506 

263.  Live-Steam  Heaters  and  Purifiers 507 

264.  Economizers 508 

265.  Value  of  Economizers 511 

266.  Factors  Determining  Installation  of  Economizers 512 

267.  Feed-Water  Temperature  due  to  Use  of  Economizers 512 

268.  Choice  of  Feed- Water  Heating  Systems 516 

CHAPTER   XIII.  —  PUMPS 522-574 

269.  Classification  of  Pumps 522 

270.  Boiler-Feed  Pumps,  Direct-Acting  Duplex 524 

271.  Boiler-Feed  Pumps,  Direct- Acting,  Steam-Actuated  Gear 527 

272.  Air  and  Vacuum  Chambers 529 

273.  Water  Pistons  and  Plungers 530 

274.  Performance  of  Piston  Pumps 531 

275.  Size  of  Boiler-Feed  Pump 539 

276.  Steam-Pump  Governors 541 

277.  Feed-Water  Regulators,  Steam  Pumps 541 

278.  Power  Pumps 543 

279.  Injectors 545 

280.  Positive  Injectors 547 

281.  Automatic  Injectors 547 

282.  Performance  of  Injectors 548 

283.  Injector  vs.  Steam  Pump  as  a  Boiler  Feeder 548 

284.  Air  Pumps 552 

285.  Dean  Wet-Air  Pump 552 

286.  Size  of  Wet-Air  Pumps  for  Jet  Condensers 553 

287.  Edwards  Air  Pump 555 

288.  Mullan  Valveless  Air  Pump 556 

289.  Alberger  Rotative  Dry-Air  Pump 557 

290.  Size  of  Wet- Air  Pumps  for  Surface  Condensers 558 

291.  Size  of  Dry- Air  Pumps  for  Surface  Condensers 558 

292.  Centrifugal  Pumps 560 

292a.   Hot-Well  Pumps 560 

293.  Volute  Centrifugal  Pumps 561 

294.  Turbine  Centrifugal  Pumps 561 

295.  Performance  of  Centrifugal  Pumps 563 

296.  Rotary  Pumps 567 

297.  Circulating  Pumps 572 

298.  Air  Lift 572 

CHAPTER   XIV.  —  SEPARATORS,  TRAPS,  AND  DRAINS 575-605 

299.  Live-Steam  Separators 575 

300.  Classification  of  Separators 576 

301.  Reverse-Current  Steam  Separators 577 

302.  Centrifugal  Steam  Separators 578 

303.  Baffle-Plate  Steam  Separators 579 

304.  Mesh  Steam  Separators 580 

305.  Location  of  Separators 580 

306.  Exhaust-Steam  Separators  and  Oil  Eliminators '.  . .  .    581 

307.  Exhaust  Heads .   585 


xii  CONTENTS 

CHAPTER   XIV  —  Continued  PAGE 

308.  Drips . .  . 586 

309.  Low-Pressure  Drips 586 

310.  Size  of  Pipes  for  Low-Pressure  Drips 588 

311.  High-Pressure  Drips 588 

312.  Classification  of  Traps 588 

313.  Float  Traps 589 

314.  Bucket  Traps 590 

315.  Dump  or  Bowl  Traps 591 

316.  Expansion  Traps 592 

317.  Differential  Traps 594 

318.  Location  of  Traps 596 

319.  Drips  under  Vacuum 597 

320.  Drips  under  Alternate  Pressure  and  Vacuum . 599 

321.  The  Steam  Loop 600 

322.  The  Holly  Loop 602 

323.  Returns  Tank  and  Pump 602 

324.  Office  Building  Drains 603 

CHAPTER   XV.  —  PIPING  AND  PIPE  FITTINGS 606-668 

325.  General 606 

326.  Drawings 606 

327.  Material  for  Pipes  and  Fittings 606 

328.  Size  and  Strength  of  Commercial  Pipe 608 

329.  Screwed  Fittings 610 

330.  Flanged  Fittings 610 

331.  Pipe  Coverings 616 

332.  Expansion  due  to  Temperature  Variation 618 

333.  Pipe  Supports  and  Anchors 621 

334.  General  Arrangement  of  High-Pressure  Steam  Piping 622 

335.  Main  Steam  Headers 629 

336.  Flow  of  Steam  in  Pipes 632 

337.  Equation  of  Pipes 636 

338.  Friction  through  Valves  and  Fittings 639 

339.  Exhaust  Piping,  Condensing  Plants 642 

340.  Exhaust  Piping,  Non-Condensing  Plant,  Webster  Vacuum  Heating 

System 642 

341 .  Exhaust  Piping,  Non-Condensing  Plant,  Paul  Vacuum  Heating  System  643 

342.  Automatic  Temperature  Control 646 

343.  Feed-Water  Piping 647 

344.  Flow  of  Water  through  Orifices,  Nozzles  and  Pipes 650 

345.  Stop  Valves 655 

346.  Automatic  Non-Return  Valves 658 

347.  Emergency  Valves 658 

348.  Check  Valves 660 

349.  Blow-off  Cocks  and  Valves 661 

350.  Safety  Valves 663 

351.  Back-Pressure  and  Atmospheric  Relief  Valves 665 

352.  Reducing  Valves;  Pressure  Regulators 666 

353.  Foot  Valves .668 


CONTENTS  xiii 

PAGE 
CHAPTER   XVI.  —  LUBRICANTS  AND  LUBRICATION 669-689 

354.  General 669 

355.  Vegetable  Oils 669 

356.  Animal  Oils  and  Fats 669 

357.  Mineral  Oils 670 

358.  Solid  Lubricants 671 

359.  Greases 671 

360.  Qualifications  of  Good  Lubricants 671 

361.  Identification  of  Oils 672 

362.  Gravity 672 

363.  Viscosity 673 

364.  Flash  Point 673 

365.  Burning  Point  or  Fire  Test 674 

366.  Acidity 674 

367.  Cold  Test 674 

368.  Friction  Test 674 

369.  Atmospheric  Surface  Lubrication 675 

370.  Intermittent  Feed 675 

371.  Restricted  Feed 675 

372.  Oil  Bath 675 

373.  Oil  Cups 677 

374.  Telescopic  Oiler 677 

375.  Ring  Oiler 678 

376.  Centrifugal  Oiler 678 

377.  Pendulum  Oiler 679 

378.  "Splash  "  Oiling 679 

379.  Gravity  Oil  Feed 680 

380.  Low-Pressure  Gravity  System 680 

381.  Compressed- Air  Feed 680 

382.  Cylinder  Lubrication 682 

383.  Cylinder  Cups 682 

384.  Hydrostatic  Lubricator 683 

385.  Forced-Feed  Cylinder  Lubricator 684 

386.  Siegrist  System  of  Lubrication 685 

387.  Oil  Filters 687 

CHAPTER   XVII.  —  FINANCE  AND  ECONOMICS  —  COST  OF  POWER 690-729 

388.  Records 690 

389.  Output 690 

390.  Load  Factor 691 

391.  Cost  of  Operation 693 

392.  Fixed  Charges 693 

393.  Interest 693 

394.  Depreciation 694 

395.  Maintenance 699 

396.  Taxes  and  Insurance 699 

397.  Operating  Costs 699 

398.  Labor,  Attendance,  Wages 699 

399.  Fuel 700 

400.  Oil,  Waste,  and  Supplies 703 

401 .  Repairs  and  Maintenance 703 

402.  Cost  of  Power .  703 


xiv  CONTENTS 

PAGE 
CHAPTER  XVIII.  —  TESTING  AND  MEASURING  INSTRUMENTS 730-749 

403.  General 730 

404.  Weighing  Fuel 730 

405.  Measurement  of  Water 730 

406.  Steam  Meters 734 

406a.   Pressure  Gauges 735 

407.  Temperature  Measurements 736 

408.  Power  Measurements 741 

409.  Flue-Gas  Analysis 741 

410.  Measurement  of  Moisture  in  Steam 745 

411.  Fuel  Calorimeters 747 

411a.   Hamler-Eddy  Smoke  Recorder 749 

CHAPTER   XIX.  —  TYPICAL  SPECIFICATIONS 750-773 

412.  Specifications  for  a  Cross  Compound  Non-Condensing  Engine 750 

413.  Specifications  for  a  Return  Tubular  Boiler 754 

414.  Specifications  for  a  Condenser  Plant 758 

415.  Specifications  for  a  Piping  System 760 

416.  Government  Specifications  for  Purchasing  Coal 769 

CHAPTER  XX.  —  A  TYPICAL  STEAM  TURBINE  STATION  —  COMMONWEALTH 

EDISON  COMPANY,  CHICAGO 774-787 

CHAPTER  XXI.  —  A  TYPICAL  ISOLATED  STATION  —  WEST  ALBANY  POWER 
STATION  OF  THE  NEW  YORK  CENTRAL  RAILROAD  COMPANY,  WEST 
ALBANY,  N.  Y 788-797 

APPENDIX    A.  —  GENERAL    BIBLIOGRAPHY  —  POWER  PLANT  ENGINEERING 

AND  DESIGN 798-821 

APPENDIX  B.  —  A.  S.  M.  E.  RULES  FOR  CONDUCTING  BOILER  TRIALS,  CODE 

OF  1899 822-845 

APPENDIX  C.  —  A.    S.    M.    E.    RULES   FOR    CONDUCTING    STEAM   ENGINE 

TESTS 846-872 

APPENDIX  D.  —  STEAM  TABLES 873-876 

APPENDIX   E.  —  EQUIVALENT   VALUES   OF   MECHANICAL   AND   ELECTRICAL 

UNITS 877 

APPENDIX  F.  —  MISCELLANEOUS  CONVERSION  TABLES 878 

APPENDIX  G.  —  RULES  FOR  FIREMEN  USING  ILLINOIS  AND  INDIANA  COAL 

IN  HAND-FIRED  FURNACES 879-880 

APPENDIX  H.  —  MOLLIER  DIAGRAM 881-885 


LIST  OF  TABLES 


PAGE 

1.  Composition  of  Typical  American  Coals •  •  •     23 

2.  Data  Relative  to  Elements  Most  Commonly  Met  with  in  Connection  with 

Combustion * 

3.  Weight  of  Air  per  Pound  of  Combustible  as  Indicated  by  the  Percentage  of 

CO2  in  the  Flue  Gas 31 

4.  Heat  Carried  away  by  the  Dry  Chimney  Gases  per  Pound  of  Combustible..     34 

5.  Loss  Due  to  Incomplete  Combustion  of  Carbon  to  Carbon  Monoxide 35 

6.  Effect  of  Washing  on  Bituminous  Coals 41 

7    Comparative  Tests  of  Babcock  &  Wilcox  Boiler.     Lump  Coal  vs.  Powdered 

Coal 46 

8.  Analyses  of  Typical  American  Fuel  Oils 52 

9.  Boiler  Efficiencies,  Fuel  Oil 54 

10.  Tests  of  Fuel-Oil  Burners 62 

11.  Characteristics  of  Gaseous  Fuels 67 

12.  Required  Hourly  Evaporation  per  Boiler  Horse  Power  at  Various  Feed 

Temperatures  and  Steam  Pressures 96 

13.  Ratio  of  Heating  Surface  to  Grate  Surface  in  Recent  Boiler  Installations 98 

14.  Examples  of  Steam  Boiler  Tests 10° 

15.  Cost  of  Evaporating  Water,  Results  of  Actual  Tests 105 

16.  Air  Spaces  and  Thickness  of  Grate  Bars 1 

17.  Values  of  cp  at  Atmospheric  Pressure  by  Various  Authorities 157 

18.  Average  Yearly  Expense  for  Repairs  for  Cast-Iron  Superheaters 171 

19.  Difference  in  Heat  Efficiency  of  Superheaters  Installed  in  Flue  and  Sepa- 

rately Fired  Superheaters • 1?6 

20.  Decrease  in  Temperature  of  Gases  of  Combustion  due  to  Superheater 

Installed  in  Flue 177 

21.  Increase  in  Heat  Efficiency  of  the  Boiler  due  to  Superheater 178 

22.  Comparative  Boiler  Tests,   Saturated  vs.  Superheated  Steam,  at  Spring 

Creek  Pumping  Station  of  the  Brooklyn  Waterworks 179 

23.  Density  and  Weight  of  Air  and  Chimney  Gas  at  Various  Temperatures 209 

24.  Theoretical  Draft  Pressures  in  Inches  of  Water  for  Various  Chimney  Tem- 

peratures     210 

25.  Test  of  a  100-Foot  Steel  Chimney -213 

26.  Chimney  Formulas 215 

27.  Size  of  Chimneys  for  Steam  Boilers 216 

28.  Approximate  Weight  and  Cost  of  Guyed  Steel  Stacks 219 

29.  Steel  Stack  Dimensions 222 

30.  Dimensions  of  Steel  Chimney  Foundations 242 

31.  Dimensions  of  Brick  Factory  Chimneys 244 

32-33.  Test  of  Steam  Jet  Blowers -248 

34.  Sizes  of  Forced-Draft  Fans 261 

35.  Sizes  of  Induced-Draft  Fans 262 

xv 


xvi  LIST  OF  TABLES 

PAGE 

36.  Steam  Engine  Efficiencies 274 

37.  Mechanical  Efficiencies  of  Engines 276 

38.  Distribution  of  Friction  Losses  in  Engines 285 

39.  Performance  of  High-Speed  Engines 296 

40.  Performance  of  Saturated-Steam  Engines,  Compound 306 

41.  Effect  of  Condensing  on  Engine  Economy , 310 

42.  Per  Cent  Moisture  Evaporated  by  Throttling 312 

43.  Performance  of  Superheated-Steam  Engines 315 

44.  Effect  of  Superheat  on  Simple  Engines 316 

45.  Effect  of  Superheat  on  Compound  Engines 317 

46.  Effect  of  Superheat  on  Triple-Expansion  Engines 318 

47.  Record  Steam  Engine  Performance.     Superheated  Steam 321 

48.  Performance  of  Steam  Turbines 390 

49.  Pressures  of  Aqueous  Vapor,  Regnault 399 

50.  Ratio  by  Weight  of  Cooling  Water  to  Steam  Condensed 406 

51.  Size  of  Siphon  Condensers 409 

52.  Square  Feet  Cooling  Surface  Necessary  to  Condense  One  Pound  of  Steam 

under  Different  Conditions 428 

53.  Test  of  a  Fennel  Saturated- Air  Condenser 432 

54.  Test  of  a  Cast-Iron  Evaporative  Surface  Condenser 434 

55.  Power  Consumption  of  Condenser  Auxiliaries '.  .   449 

56.  Most  Economical  Vacuum  for  Steam  Turbines 452 

57.  Most  Economical  Vacuum  for  Piston  Engines 453 

58.  Properties  of  Saturated  Air 463 

59.  Influence  of  Thickness  of  Scale  on  Heat  Transmission 472 

60.  Water  and  Boiler  Scale  Analyses 473 

61.  Boiler  Defects,  Report  of  Hartford  Steam  Boiler  Inspection  and  Insurance 

Company.  . 474 

62.  Percentage  of  Saving  for  each  Degree  of  Increase  in  Temperature  of  Feed 

Water 485 

63.  Feed- Water  Temperatures,  Open  Heaters 490 

64.  Extent  of  Heating  Surface,  Closed  Heaters 498 

65.  Mean  Temperature  Difference,  Closed  Heaters 499 

66.  Heat  Transmission,  Closed  Heaters 502 

67.  Economizer  Tests 515 

68.  Pump  Duties  for  Various  Efficiencies  and  Steam  Consumptions 537 

69.  Maximum  Height  to  which  Pumps  can  Raise  Water  by  Suction  —  Tem- 

perature Constant 538 

70.  Maximum    Height    to    which    Pumps    can    Raise    Water  —  Temperature 

Variable 539 

71.  Range  of  Working  Pressure  —  Metropolitan  Injectors 550 

72.  Commercial  Sizes  of  Air  Pumps  for  Condensers 558 

73.  Data  Pertaining  to  Single-Stage  Centrifugal  Pumps 566 

74.  Data  Pertaining  to  Multi-Stage  Centrifugal  Pumps 567 

75.  Tests  of  Steam  Separators 576 

76.  Dimensions  of  Standard  Wrought-!  ron  Pipes 611 

77.  Comparative  Costs  of  Different  Types  of  Flanges 614 

78.  Dimensions  of  Standard  Flanges 614 

79.  Dimensions  of  Extra  Heavy  Flanges 615 

80.  Loss  of  Heat  from  Bare  Pipes  in  Still  Air 615 

81.  Experiments  on  Pipe  Coverings .  .  ; 617 


LIST  OF  TABLES  xvii 

PAGE 

82.  Coefficients  of  Expansion  —  Piping  Materials 620 

83.  Comparison  of  Formulas  for  the  Flow  of  Steam  in  Pipes 634 

84.  Comparison  of  Formulas  for  the  Flow  of  Steam  in  Pipes 635 

85.  Flow  of  Steam  in  Pipes,  Babcock 637 

86.  Flow  of  Steam  in  Pipes,  Sickles 638 

87.  Equation  of  Pipes 640 

88.  Specific  Gravity  of  Lubricating  Oils 673 

89.  Properties  of  Lubricating  Oils 670,  676 

90.  Approximate   Useful  Life  of  Various  Portions  of  Steam  Power  Plant 

Equipments 694 

91.  Rates  of  Depreciation 695 

92.  Depreciation  Percentages  —  Chicago  Traction  Valuation  Commission 696 

93.  Cost  of  Labor  for  Street  Railway  Plants 701 

94.  Cost  of  Labor  for  Tall  Office  Buildings 702 

95.  Operating  Costs  per  Kilowatt  Hour,  Typical  British  Electric  Light  and 

Power  Plants 709 

96.  Operating  Costs  per  Kilowatt  Hour,  Typical  United  States  Railway  Plants  709 

97.  Operating  Costs  per  Kilowatt  Hour,  Average  of  all  Stations,  Boston  Ele- 

vated     710 

98.  Operating  Costs  (1907),  First  National  Bank  Building,  Chicago 710 

99.  Cost  of  One  Horse  Power  per  Year,  Simple  Engine,  W.  O.  Webber 711 

100.  Cost  of  One  Horse  Power  per  Year,  Compound  Engine,  W.  O.  Webber.  .  .   711 

101.  Cost  of  One  Horse  Power  per  Year,  H.  von  Schon 712 

102.  Cost  of  Electrical  Power  per  Year,  W.  M.  Wilson 713 

103.  Cost  of  Electrical  Power  per  Year,  R.  C.  Carpenter 715 

104.  Cost  of  Electrical  Power  per  Year,  Oil  Fuel,  C.  C.  Moore  &  Co 716 

105-7.   Cost  of  Power,  Typical  Isolated  Stations 720,  722 

108.  Cost  of  Power,  First  National  Bank  Building,  Chicago 724 

109.  Temperature  Ranges  of  Thermometers  in  General  Use 739 

Appendix  D.  —  Steam  Tables 873 

ADDITIONAL  TABLES — THIRD  EDITION. 

00.   Physical  and  Chemical  Properties  of  Woods,  Straw  and  Tanbark 19 

0.   Heat  Values  of  Bagasse  and  Variation  with  Degree  of  Extraction 20 

2a.   Ratio  of  Total  Air  Supplied  to  that  Theoretically  Required  for  Various 

Analyses  of  Flue  Gases 30 

14a.   Principal  Data  and  Results,  Boiler  Unit  No.  10,  Fisk  Street  Station,  Com- 
monwealth Edison  Company 102 

15a.    Pounds  of  Water  Evaporated  per  Hour  from  and  at  212°  F.  per  Pound  of 

Fuel 108 

17a.   Mean  Specific  Heat  of  Superheated  Steam 159 

17b.   Specific  Volume  of  Superheated  Steam 160 

94a.   Distribution  of  Maintenance  and  Operating  Costs  in  Large  Stations 707 

107a.    Yearly  Operating  Costs  in  Four  Typical  Central  Stations 723 

107b,   Power  Costs  —  Steam  Electric  Central  Stations. .  .  723 


ILLUSTRATIONS. 


CHAPTER  I 

ELEMENTARY  STEAM  POWER  PLANTS. 
FIG. 

1.  Elementary  Non-Condensing  Plant 

2.  Elementary  Non-Condensing  Plant  with  Heating  System. 

3.  Simple  Condensing  Plant. 

4.  Condensing  Plant  with  Full  Complement  of  Heat-Saving  Appliances. 
4a.  Sectional  Elevation  of  the  Myers  Furnace  for  Burning  Bagasse. 

4b.  End  and  Side  Sectional  Elevation  of  the  Myers  Furnace  for  Burning  Tanbark. 

CHAPTER  II.  —  FUELS  AND  COMBUSTION. 

5.  Relation  of  Gas  Composition  in  Combustion  Chamber  to  Temperature. 

6.  Influence  of  Size  of  Coal  on  Boiler  Capacity  and  Efficiency. 

7.  Influence  of  Ash  on  Fuel  Value  of  Dry  Coal. 

8.  Pinther  Coal-dust  Feeder. 

9.  Schwartzkopff  Coal-dust  Feeder. 

10.  Aero-Pulverizer  Coal-dust  Feeder. 

11.  Triumph  Coal-dust  Feeder. 

12.  Korting  Fuel  Oil  Burner. 

13.  Booth  Fuel  Oil  Burner. 

14.  Hammel  Fuel  Oil  Burner. 

15.  Branch  Fuel  Oil  Burner. 

16.  Kirkwood  Fuel  Oil  Burner. 

17.  Williams  Fuel  Oil  Burner. 

18.  Warren  Fuel  Oil  Burner. 

19.  Furnace  for  Burning  Fuel  Oil,  Front  Feed. 

20.  Furnace  for  Burning  Fuel  Oil,  Rear  Feed. 

21.  International  Gas  and  Fuel  Company's  Fuel  Oil  System. 

22.  Hydraulic  Oil  Storage  Company's  Fuel  Oil  System. 

CHAPTER  III.  —  BOILERS. 

23.  Vertical  Tubular  Boiler  with  Submerged  Tube  Sheet. 

24.  Manning  Vertical  Fire-Tube  Boiler. 

25.  Typical  Fire-box  Boiler,  —  Stationary  Type. 

26.  Stationary  Scotch-Marine  Boiler. 

27.  Robb-Mumford  Boiler. 

28.  Return  Tubular  Boiler  Setting,  —  Extended  Front. 

29.  Return  Tubular  Boiler  Setting,  —  Flush  Front. 

30.  Return  Tubular  Boiler  Setting,  —  Steel  Beam  Suspension. 

31.  Boiler  Setting,  "Wood"  Mill  of  the  American  Woolen  Company,  Lawrence, 

Mass. 

32.  Furnace  Arch  Bars. 

33.  Back  Connection  made  with  Cast-Iron  Plate. 

34.  Babcock  and  Wilcox  Boiler  and  Setting. 

xix 


xx  ILLUSTRATIONS 

FIG. 

35.  Details  of  Header,  —  Babcock  and  Wilcox  Boiler. 

36.  Front  Section,  —  Babcock  and  Wilcox  Boiler. 

37.  Heine  Boiler  and  Setting. 

38.  Wickes  Vertical  Water-Tube  Boiler. 
38a.    1200-H.P.  Parker  Boiler. 

38b.    Boiler  Room  Area  for  Various  Types  of  Boilers. 

39.  Stirling  Boiler  and  Setting. 

39a.   Heat  Transmission  Through  Boiler  Plate. 

39b.    Influence  of  Draft  on  Capacity,  Torpedo  Boat  "Biddle." 

40.  Influence  of    Draft    on  the  Efficiency  and    Capacity  of    a  350-Horse-power 

Babcock  and  Wilcox  Boiler  with  Chain  Grate. 

41.  Effect  of  Rate  of  Driving  on  Economy  of  a    150-Horse-power  Stirling  Boiler, 

Hand  Fired. 
41a.    Relation  Between  Efficiency  and    Capacity,   500-Horse-power  Babcock  and 

Wilcox  Boiler. 
41b.    Effect  of  Rate  of  Driving  on  Efficiency  of  a  600-Horse-power  Babcock  and 

Wilcox  Boiler. 
41c.    Influence  of  Draft  on  the  Capacity  of  a  600-Horse-power  Babcock  and  Wilcox 

Boiler. 

42.  Effect  of  Thickness  of  Fire  on  the  Capacity  and  Efficiency  of  a  350-Horse- 

power  Stirling  Boiler,  equipped  with  Chain  Grate. 

43.  Effect  of  Thickness  of  Fire  on  the  Capacity  and  Efficiency  of  a  150-Horse- 

power  Water-Tube  Boiler. 

44.  Effect  of  Thickness  of  Fire  on  the  Capacity  and  Efficiency  of  a  500-Horse- 

power  Babcock  and  Wilcox  Boiler. 

45.  Types  of  Grate  Bars. 

46.  A  Typical  Rocking  Grate. 

47.  Horizontal  Blow-off  Connection  to  Head. 

48.  Vertical  Blow-off  Connection  to  Shell. 

49.  Blow-off  Connection  with  Circulating  Pipe. 

50.  Blow-off  Tank  and  Connections. 

51.  Surface  Blow-off. 

52.  Buckeye  Skimmer. 

53.  Kitts  Hydraulic  Damper  Regulator. 

54.  Tilden  Steam  Actuated  Damper  Regulator. 

55.  Simple  Water  Column. 

56.  Water  Gauge  with  Self-closing  Valve. 

57.  Combined  Water  Column  and  High  and  Low  Water  Alarm. 

58.  Types  of  Fusible  Plugs. 

59.  Mechanical  Tube  Cleaner,  —  Hammer  Type. 

60.  Mechanical  Tube  Cleaner,  —  Turbine  Type. 

CHAPTER  IV.  —  SMOKE  PREVENTION,  FURNACES,  STOKERS. 

61.  Green  Chain  Grate. 

€2.  Babcock  and  Wilcox  Boiler,  Chain  Grate,  Ordinary  Setting. 

63.  Babcock  and  Wilcox  Boiler,  Chain  Grate,  Fire-tile  Roof. 

64.  Section  of  Fire  Tile. 

65.  Section  of  Fire  Tile. 

66.  Application  of  "Economy"  Fire  Files  to  Stirling  Boiler. 

67.  Method  of  Anchoring  "Economy"  Fire  Tiles  to  Tubes. 
67a.   Chain  Grate  Fired  from  Rear  End  of  Setting. 


ILLUSTRATIONS  xxi 

FIG. 

67b.    Smokeless  Setting,  Chain  Grate  and  Babcock  and  Wilcox  Boiler. 

68.  Details  of  Roney  Stoker. 
68a.    Double  Stoker  Setting. 

69.  Details  of  Wilkinson  Stoker. 

70.  Murphy  Furnace,  Front  Section. 

71.  Murphy  Furnace,  Side  Section. 

72.  Jones  Underfeed  Stoker. 

73.  American  Underfeed  Stoker. 

74.  Hawley  Down-Draft  Furnace. 

75.  Plain  Dutch  Oven. 

76.  "Twin  Fire  Arch,"  Applied  to  a  Return-Tubular  Boiler. 

76  a,  b,  c.    Chicago  Setting,  Hand-Fired  Return-Tubular  Boiler. 

77.  Wooley  Smokeless  Furnace. 

78.  Kent's  Wing-Wall  Furnace. 

79.  Burke's  Smokeless  Furnace,  Front  Section. 

80.  Burke 's  Smokeless  Furnace,  Side  Section. 

81.  Split  Bridge  Wall. 

CHAPTER  V.  —  SUPERHEATED  STEAM;  SUPERHEATERS. 

82.  Specific  Heat  of  Superheated  Steam,  Knoblauch  and  Linde. 

83.  Specific  Heat  of  Superheated  Steam,  A.  R.  Dodge. 

84.  Specific  Heat  of  Superheated  Steam,  C.  E.  Burgeon. 

85.  Specific  Heat  of  Superheated  Steam,  Thomas. 

86.  Babcock  and  Wilcox  Superheater. 

87.  Stirling  Superheater. 

88.  Details  of  Stirling  Superheater. 

89.  Foster  Superheater  in  Babcock  and  Wilcox  Boiler. 

90.  Schmidt  Independently  Fired  Superheater. 

91.  Foster  Independently  Fired  Superheater. 

92.  Schmidt    System   of    Combined    Superheater,   Economizer   and    Feed-Water 

Heater. 

92a.    Relation  between  Gas  Temperature,  Heating  Surface  passed  over  and  Amount 
of  Steam  Generated. 

93.  Relation  of  Superheat  to  Total  Output  of  Boiler. 

94.  Relation  of  Output  of  Superheater  to  Boiler  Output. 

95.  Relation  of  Superheat  to  Output  of  Superheater. 

CHAPTER  VI.  —  COAL  AND  ASH  HANDLING  APPARATUS. 

96.  Link-Belt  Coal-Handling  Apparatus. 

97.  Typical  Coal  and  Ash  Handling  Equipment. 

98.  Steel  Cable  Company's  Coal-Handling  Apparatus. 

99.  Coal  and  Ash  Handling  System  of  S.  S.  Elevated  Railway. 

100.  Crusher  and  Conveyor  of  S.  S.  Elevated  Railway. 

101.  Driving  Mechanism  of  Hunt  Conveyor. 

102.  Hunt  Coal  Conveyor  System  at  Baltimore,  Md. 

103.  Bucket  and  Screw  Conveyor  at  Commercial  National  Bank  Building,  Chicago,, 

Illinois. 

104.  Guide  Pulleys,  Robins  Belt  Conveyor. 

105.  Coal  and  Ash  Handling  System  of  Aurora  and  Elgin  Railway. 

106.  Coa]  and  Ash  Handling  System  of  Cincinnati  Traction  Company. 


xxii  ILLUSTRATIONS 

PIG. 

107.  Coal  and  Ash  Handling  System  of  Detroit  Edison  Company. 

108.  Vacuum  Ash  Handling  System. 

108a.    Vacuum  Ash  Handling  System  at  the  Armour  Glue  Works. 
108b.   Coal  and  Ash  Handling  System,  Norfolk  Traction  Company. 

109.  Stationary  Coal- Weighing  Hoppers. 

110.  Traveling  Coal  Hoppers. 

111.  Common  Slide  Coal  Valve. 

112.  Simplex  Coal  Valve. 

113.  Duplex  Coal  Valve. 

114.  Flap  Coal  Valve. 

115.  Seaton  Coal  Valve. 

CHAPTER  VII.  —  CHIMNEYS. 

116.  Relation  between  Draft  and  Rates  of  Combustion. 

117.  Steel  Chimney  of  S.  S.  Elevated  Railway  Power  House,  Chicago. 

118.  Stability  of  Steel  Chimneys. 

119.  Custodis  Radial  Brick  Chimney. 
119a.   Custodis  Radial  Perforated  Brick. 

120.  Circular  Brick  Chimney  at  Armour  Institute  of  Technology. 

121.  Design  of  Brick  Chimneys,  Thickness  of  Walls. 

122.  Design  of  Brick  Chimneys,  Stability. 

123.  Weber  Steel-Concrete  Chimney. 

124.  Weber  Steel-Concrete  Chimney. 

CHAPTER  VIII.  —  MECHANICAL  DRAFT. 

125.  Ring  Steam  Jet. 

126.  Bloomsburg  Jet. 

127.  McClaves  Argand  Blower. 

128.  Hollow  Bridge  Wall  and  Steam  Jet. 

129.  Parsons  Smokeless  Furnace. 

130.  Heinrich  Smokeless  Furnace.     (Sectional  Elevation.) 

131.  Heinrich  Smokeless  Furnace.     (Sectional  Plan.) 

132.  Typical  Forced-Draft  System. 

133.  Typical  Induced-Draft  System. 

134.  Pitot  Tubes;  Orifice  Closed. 

135.  Pitot  Tubes;  Orifice  Wide  Open. 

136.  Pitot  Tubes;  Orifice  Partly  Closed. 

137.  Performance  of  Steel  Plate  Fans. 

138.  Performance  of  Pressure  Blower,  Speed  Constant. 

139.  Performance  of  Pressure  Blower,  Speed  Variable. 

140.  Comparative  Costs  of  Chimneys  and  Mechanical  Draft. 

141.  Influence  of  Rate  of  Combustion  on  Air  Supply;  Forced  Draft. 

142.  Balanced  Draft  System. 

CHAPTER  IX.  —  STEAM  ENGINES. 
142a.    Rankine  Cycle. 

142b.    Side  Elevation,  Typical  Corliss  Engine. 
142c.    Plan  View,  Typical  Corliss  Engine. 
142d.    A  Modern  Piston  Engine  Plant. 

143.  Mechanical  Efficiencies  of  Engine  and  Generator. 

144.  Status  of  the  Steam  Engine. 

145.  Condensation  and  Leakage  Losses  in  Simple  Engines. 


ILLUSTRATIONS  xxiii 

FIG. 

145a.    Influence  of  Increasing  Back  Pressure. 

146.  Typical  Curves  of  Steam  Engine  Friction. 

147.  Influence  of  Increasing  Initial  Pressure. 

148.  Typical  Economy  Curves  —  High-Speed  Engines. 
148a.    Assembly  of  Valve  Gear,  Typical  Corliss  Engine. 
148b.    Section  through  Cylinder,  Typical  Corliss  Engine. 

148c.    Assembly  of  Governor  and  Link  Mechanism,  Corliss  Engine. 

149.  Test  of  Reeves  Simple  Engine;  Condensing  vs.  Non-Condensing. 

150.  Typical  Economy  Curves  of  Single- Valve  vs.  Four- Valve  High-Speed  Engines. 
150a.   3500-K.W.  Vertical  Cross  Compound  Corliss  Engine. 

150b.    7500-K.W.  Vertical  Horizontal  Cross  Compound  Corliss  Engine. 

151.  Effect  of  Compounding  on  High-Speed  Non-Condensing  Engines. 

152.  Performance  of  Corliss  Compound;  Condensing  vs.  Non-Condensing. 

153.  Performance  of  a  5500-Horse-power  Engine. 

154.  Increase  in  Power  Due  to  Vacuum. 

155.  Increase  in  Power  Due  to  Vacuum. 

156.  Performance  of  a  5500-Horse-power  Engine. 

157.  Indicator  Cards  —  High-Speed  Throttling  Engines. 

158.  Indicator  Cards  —  High-Speed  Automatic  Engines. 

159.  Effect  of  Superheat  on  Steam  Consumption. 

159a.   3000-Horse-power  Sulzer  Engine  for  Highly  Superheated  Steam. 
159b.    Fleming-Harrisburg  Four- Valve  Tandem  Compound. 

160.  Effect  of  Superheat  on  Steam  Compounds. 

161.  Influence  of  Superheat  on  Economy. 

162.  Diagrammatic  Arrangement,  Binary- Vapor  Engine. 

163.  Cost  of  Simple  High-Speed  Engines. 

164.  Cost  of  High-Speed  Compound  Engines. 

165.  Cost  of  Simple  and  Low-Speed  Compound  Engines. 

CHAPTER  X.  —  STEAM  TURBINE. 

166.  Horizontal  Section  of  De  Laval  Turbine. 

167.  Details  of  Blades  of  De  Laval  Turbine. 

168.  Details  of  Nozzle  of  De  Laval  Turbine. 

169.  Details  of  Governor  of  De  Laval  Turbine. 

170.  Theoretically  Proportional  Expanding  Nozzle. 

171.  Theoretical  Performance  of  a  Divergent  Nozzle. 

172.  Characteristic  Performance  of  a  Divergent  Nozzle. 
172a.    Velocity  Diagram,  Ideal  Impulse  Turbine. 

172b.    Velocity  Diagram,  as  Modified  by  Friction  Losses. 

173.  Section  through  Terry  Turbine. 

174.  Arrangement  of  Buckets  and  Reversing  Chambers,  Terry  Turbine. 

175.  Longitudinal  Section  through  Kerr  Turbine. 

176.  Sectional  End  Elevation,  Kerr  Turbine. 

177.  Details  of  Governor,  Kerr  Turbine. 

178.  Four-Stage  Vertical  Curtis  Turbo-Generator. 

179.  3500-K.W.  Horizontal  Curtis  Turbine. 

180.  Arrangement  of  Nozzles  and  Blades,  Curtis  Turbine. 

181.  Section  through  Curtis  Governor. 

182.  Mechanical  Valve  Gear,  Curtis  Turbine. 

183.  Hydraulic  Valve  Gear,  Curtis  Turbine. 

183a.    Steam  Belt  Area  in  Five-Stage  Curtis  Turbine. 


xxiv  ILLUSTRATIONS 

FIG. 

184.  Velocity  Diagram,  Curtis  Turbine. 

185.  Section  through  Hamilton-Holzworth  Turbine. 

186.  Details  of  Vanes,  Hamilton-Holzworth  Turbine. 

187.  Details  of  Bearings,  Hamilton-Holzworth  Turbine 

188.  Details  of  Governor,  Hamilton-Holzworth  Turbine. 

189.  Section  through  Westinghouse-Parsons  Standard  Turbine. 

190.  Flow  of  Steam  in  Parsons  Turbine. 

191.  Details  of  Governor,  Westinghouse-Parsons  Turbine. 

192.  Indicator  Cards,  Westinghouse-Parsons  Turbine. 

193.  By-Pass  Valve,  Westinghouse-Parsons  Turbine. 

193a.  Method  of  Fastening  Blades,  Westinghouse-Parsons  Turbine. 
193b.  High-Pressure  Double-Flow,  Westinghouse-Parsons  Turbine. 
193c.  Allis-Chalmers  Steam  Turbine. 

194.  Velocity  Diagram,  Multi-Stage  Reaction  Turbine. 

194a.    Low-Pressure  Turbine,  59th  St.  Station,  Interborough  Rapid  Transit  Company. 

194b.    Low-Pressure  Double-Flow  Westinghouse-Parsons  Turbine. 

194c.    Performance  of  7500-K.W.  Engine  at  59th  St.  Station. 

194d.    Comparison  of  Economy  Curves,  Combined  Engine  and  Turbine. 

195.  Rateau  Low-Pressure  Steam  Turbine  at  South  Chicago. 

196.  Rateau  Regenerator  Accumulator. 

196a.   Typical  Double-Deck  Turbine  Installation. 

197.  Curve  of  Performance  of  Rateau  Low-Pressure  Turbine. 

198.  Comparative  Floor  Space,  Engines  vs.  Turbines. 

198a.   Typical  Performance  of  90-Horse-power  Terry  Turbine. 
198b."  Typical  Performance  9000-K.W.  Curtis  Turbine. 
198c.    Typical  Performance  Small  Non-Condensing  Turbines. 
198d.   Typical  Correction  Curves,  125-K.  W.  Turbines. 

199.  Reciprocating  Engine  vs.  Turbine  Economy. 

200.  Effect  of  Superheat  on  Economy. 

201.  Effect  of  Vacuum  on  Economy,  Westinghouse-Parsons  Turbine. 
201a.    Effect  of  a  Vacuum  on  Economy. 

202.  Effect  of  Vacuum  and  Superheat  on  Economy. 

CHAPTER  XL  —  CONDENSERS. 

203.  Worthington  Jet  Condenser. 

204.  Blake  Jet  Condenser. 

205.  Baragwanath  Siphon  Condenser. 

206.  Schutte  Ejector  Condenser. 

207.  Piping  for  Schutte  Condenser. 

208.  Weiss  Counter-Current  Condenser. 

209.  Alberger  Barometric  Condenser. 

210.  Worthington  Barometric  Condenser. 

211.  Tomlinson  Type  B  Barometric  Condenser. 

21  la.    Centrifugal  Pump  Applied  to  Tail  Pipe  of  a  Barometric  Condenser. 

212.  Baragwanath  Surface  Condenser. 

213.  Wheeler  Surface  Condenser  and  Pumps. 

214.  Wheeler  Multi-flow  Surface  Condenser. 

215.  Weighton  Multi-flow  Surface  Condenser. 

216.  Relation  between  Hot-well  Temperature  and  Vacuum  in  Surface  Condensers. 
216a.   Application  of  Weighton  Dry-Tube  Surface  Condenser  to  Vertical  Marine 

Eneine. 


ILLUSTRATIONS  xxv 

FIG. 

216b.   Heat  Transfer  in  Condenser  Tubes,  Steam  to  Water. 

216c,  d.    Heat  Transfer  in  Condenser  Tubes,  Steam  to  Air. 

217.  Fennel  Saturated-Air  Surface  Condenser. 

218.  Fennel  Flask  Type  Atmospheric  Condenser. 

219.  Jet  Condenser  located  below  Engine-Room  Floor. 

220.  Surface  Condenser  located  below  Engine-Room  Floor. 

221.  Surface  Condenser  Connected  with  Pumping  Engine. 

222.  Jet  Condenser  located  above  Engine-Room  Floor. 

223.  Typical  Arrangement,  Westinghouse-Leblanc  Condenser  and  Curtis  Turbine, 

224.  Elevation  of  Condenser  Piping,  Des  Moines  City  Railroad  Power  House. 

225.  Plan  of  the  Condenser  Piping,  Des  Moines  City  Railroad  Power  House. 

226.  Plan  of  Condenser  Piping,  Northwestern  Elevated  Railroad  Power  House, 

Chicago. 
226a.    Condenser  Installation,  Quincy  Point  Power  Plant. 

227.  Worthington  High-Vacuum  System. 

228.  Wheeler  High- Vacuum  System. 

229.  Parsons  Vacuum  Augmenter. 

229a.   Westinghouse-Leblanc  High  Vacuum  Multi-Jet  Condenser. 
229b.   Tomlinson  Type  C  High  Vacuum  Jet  Condener. 
229c.    Korting  Multi-Jet  Condenser. 

230.  Power  Consumption  of  Auxiliaries,  Parsons  Turbine. 

231.  Power  Consumption  of  Auxiliaries,  Curtis  Turbine. 

232.  Relative  Cost  of  High- Vacuum  Condensing  Systems. 

233.  Performance  of  Spray  Fountain. 

234.  Barnard-Wheeler  Cooling  Tower. 

235.  Worthington  Cooling  Tower. 

236.  Alberger  Cooling  Tower  Installation. 

CHAPTER  XII.  —  FEED-WATER  HEATERS. 

237.  Scaife  System  for  Feed-Water  Purification. 

238.  We-Fu-Go  System  for  Feed-Water  Purification. 
238a.    Anderson  System  for  Feed-Water  Purification. 

239.  Cochrane  Feed-Water  Heater  and  Receiver. 

240.  Webster  Star  Vacuum  Heater. 

241.  Hoppes  Horizontal  Heater  and  Purifier. 

242.  Goubert  Single-flow  Heater. 

243.  Expansion  Joint,  Goubert  Heater. 

244.  Wainwright  Multi-flow  Closed  Heater. 

245.  Coil  Heater. 

246.  Otis  Steam-Tube  Feed-Water  Heater. 

247.  Baragwanath  Steam-Jacketed  Heater. 

248.  Heat  Transmission  in  Feed-Water  Heater  Tubes. 

249.  Open  Heater  Connected  as  a  Through  Heater 

250.  Through  Heater  with  By-Pass. 

251.  Open  Induced  Heater,  Non-Condensing  Plant. 

252.  Closed  Induced  Heater,  Condensing  Plant. 

253.  Hoppes  Live-Steam  Heater  and  Purifier. 

254.  Installation  of  a  Live-Steam  Purifier. 

255.  Typical  Installation  of  Primary  and  Secondary  Heater. 


xxvi  ILLUSTRATIONS 

FIG. 

256.  Green  Economizer. 

257.  Economizer  Installation  at  Weehawken,  New  Jersey. 

258.  Heat  Transmission,  Economizers. 

CHAPTER  XIII.  —  PUMPS. 

259.  Duplex  Direct-Acting  Boiler-Feed  Pump. 

260.  Section  through  Duplex  Boiler-Feed  Pump. 

261.  Method  of  Obtaining  Lost  Motion,  —  Duplex  Valve  Gear. 

262.  Method  of  Obtaining  Lost  Motion,  —  Duplex  Valve  Gear. 

263.  Position  of  Valve  and  Piston  at  the  Beginning  of  Stroke. 

264.  Position  of  Valve  and  Piston  at  the  End  of  Stroke. 

265.  Pump  Disk  Valve. 

266.  Section  through  a  Compound  Duplex  Pump. 

267.  Section  through  a  Simplex  Pump  with  Steam-Actuated  Gear. 

268.  Forms  of  Vacuum  Chambers. 

269.  Different  Arrangements  of  Vacuum  Chambers. 

270.  Types  of  Water  Pistons. 

271.  Plunger  with  Metal  Packing  Ring. 

272.  Plunger  with  Hydraulic  Packing. 

273.  Horizontal  Fly-Wheel  Pump  with  Outside  Packed  Plunger. 

274.  Performance  of  Direct-Acting  Pressure  Pumps. 

275.  Performance  of  Boiler-Feed  Pump  at  the  Armour  Institute  of  Technology. 

276.  Fisher  Pump  Governor. 

277.  Kitts  Feed- Water  Regulator. 

278.  Rowe  Feed-Water  Regulator. 

279.  Triplex  Pump. 

279a.    Performance  of  Triplex  Pump,  Direct  Connected. 

280.  Performance  of  Triplex  Pump,  Geared. 

281.  Elementary  Form  of  Ejector. 

282.  The  Hancock  Inspirator. 

283.  The  Penberthy  Automatic  Injector. 

284a.    Performance  of  a  Desmond  Automatic  Injector  with  Varying  Initial  Pressure. 
284b.    Performance  of    a    Desmond    Automatic    Injector    with    Varying    Suction 

Temperature. 
284c.    Performance   of  a   Desmond   Automatic   Injector  with   Varying  Discharge 

Pressure. 

285.  Dean  Jet  Condenser  Air  Pump. 

286.  Edwards  Air  Pump. 

287.  Mullan  Valveless  Air  Pump. 

288.  Hewes  and  Phillips  Air  Pump. 

289.  Alberger  Dry-Air  Pump. 

290.  Air  Pump  Indicator  Diagram. 

291.  Types  of  Impellers,  Centrifugal  Pumps. 

292.  A  Typical  Centrifugal  Pump. 

293.  Direction  of  Water  from  Impeller  of  Volute  Pumps  without  Diffusion  Vanes. 

294.  Effect  of  Diffusion  Vanes  on  the  Direction  of  Water. 

295.  Three-Stage  Lea-Degan  Turbine  Pump. 

296.  Six-Stage  Rateau  Turbine  Pump. 

297.  Test  of  Centrifugal  Pump  at  Armour  Institute. 

298.  Centrifugal  Pump  Characteristic  for  Boiler-Feed  Pumps. 


ILLUSTRATIONS  xxvii 

FIG. 

299.  Centrifugal  Pump  Characteristic  for  Dry-Dock  Service. 

300.  Centrifugal  Pump  Characteristic  for  Waterworks  Service. 

301.  Performance  of  a  6-inch  Worthington  Conoidal  Centrifugal  Pump. 

302.  Performance  of  a  Single-Stage  De  Laval  Volute  Pump. 

303.  Performance  of  a  Two-Stage  Turbine  Pump. 

304.  Performance  of  a  Two-Stage  De  Laval  Centrifugal  Pump. 

305.  Two-Lobe  Cycloidal  Rotary  Pump. 

306.  Rotary  Pump  with  Movable  Butment. 

307.  Performance  of  a  Rotary  Pump. 

308.  High-Duty  Circulating  Pump,  New  York  Rapid  Transit  Company. 

309.  Pulsometer. 

310.  Air  Lift. 

CHAPTER  XIV.  —  SEPARATORS,  TRAPS,  DRAINS. 

311.  Hoppes  Live-Steam  Separator. 

312.  Stratton  Live-Steam  Separator. 

313.  Keystone  Live-Steam  Separator. 

314.  Bundy  Live-Steam  Separator. 

315.  Austin  Live-Steam  Separator. 

316.  Direct  Live-Steam  Separator. 

317.  Baum  Oil  Separator. 

318.  Loew  Grease  Extractor. 

319.  Typical  Exhaust  Head. 

320.  Closed  Heater  as  a  Blow-off  Tank. 

321.  Piping  Drips  to  Exhaust  Pipe. 

322.  McDaniel  Float  Trap. 

323.  Acme  Bucket  Trap. 

324.  Bundy  Tilting  Trap. 

325.  Columbia  Expansion  Trap. 

326.  Geipel  Expansion  Trap. 

327.  Dunham  Expansion  Trap. 

328.  Heintz  Expansion  Trap. 

329.  Flinn  Differential  Trap. 

330.  Simple  Siphon  Trap. 

331.  Location  of  Return  Trap. 

332-3.   Drainage  for  Jackets  and  Receivers  of  Triple  Expansion  Pumping  Engines. 

334.  Gravity  Drainage;  Vacuum  Heater. 

335.  Method  of  Draining  Heater  under  Vacuum. 

336.  Method  of  Draining  Receivers  under  Alternate  Pressure  and  Vacuum 

337.  Steam  Loop. 

338.  Holly  Loop. 

339.  Section  through  Holly  Receiver. 

340.  Returns  Tank  and  Pump. 

341.  Shone  Ejector. 

CHAPTER  XV.  —  PIPING  AND  PIPE  FITTINGS. 

342.  United  States  Standard  Pipe  Thread. 

343.  Types  of  Pipe  Flanges. 

344.  Efficiencies  of  Various  Pipe  Coverings. 

345.  Pipe  Bends. 

346.  Double-Swing  Screwed  Fittings  for  Expansion. 


xxviii  ILLUSTRATIONS 

FIG. 

347.  Slip  Expansion  Joint. 

348.  Typical  Pipe  Hanger. 

349.  Typical  Wall  Bracket  with  Roll  Binder. 
349a.   A  Typical  Floor  Stand. 

350.  Typical  Pipe  Anchor. 

351.  Arrangement  of  Steam  Piping,  Princeton  University  Power  Plant. 

352.  Typical  "Duplicate"  Header  System. 

353.  Typical  "Loop  Header"  System. 

354.  Typical  "By-Pass"  Piping  System. 

355.  General  Arrangement  of  Steam  and  Exhaust   Piping,   Heyworth  Building, 

Chicago,  Illinois. 

356.  General  Arrangement  of  Piping,  Manhattan  Elevated  Station,  New  York. 
357-8.    Piping  Arrangement  at  the  Yonkers  Power  House  of  the  New  York  Central. 

359.  Overhead  Piping  of  Boilers,  Quincy  Point  Power  Plant  of  the  Old  Colony  Street 

Railway  Company,  Quincy  Point,  Mass. 

360.  Main  Stream  Header  and  Branches,  Grand  Rapids,  Grand  Haven  and  Muskegon 

Railway  Power  House. 

361.  Main  Header  and  Branches,  Des  Moines  City  Railway  Power  House. 

362.  Drop  in  Pressure  in  Steam  Pipes  of  Various  Diameters  at  Different  Velocities. 

363.  Diagrammatic  Arrangement  of  Piping  in  the  Webster  Vacuum  Heating  System. 

364.  Webster  Vacuum  Seal  Valve. 

365.  Automatic  Vacuum  Valve,  Illinois  Engineering  Company. 

366.  Diagrammatic  Arrangement  of  Piping  in  the  Paul  Vacuum  Heating  System. 

367.  Paul  Exhauster. 

368.  Paul  Vacuum  Valve. 

369.  Powers  Thermostat. 

370.  Typical  Diaphragm  Valve. 

371.  Diagram  of  Feed- Water  Piping,  Condensing  Plant. 

372.  Diagram  of  Feed- Water  Piping,  Non-Condensing  Plant. 

373.  Arrangement  of  Valves  in  Feed-Water  Branches. 

374.  Globe  Valve,  Screw-Top,  Inside  Screw. 

375.  Globe  Valve,  Bolt-Top,  Outside  Screw. 

376.  Gate  Valve,  Solid- Wedge,  Screw-Top,  Outside  Screw. 

377.  Gate  Valve,  Solid-Wedge,  Bolt-Top,  Inside  Screw. 

378.  Gate  Valve,  Split-Wedge,  Bolt-Top,  Inside  Screw. 

379.  Ludlow  Angle  Valve,  Gate  Pattern. 

380.  Anderson  Automatic  Non-Return  Valve. 

381.  Crane  Hydraulic  Emergency  Gate  Valve. 

382.  Anderson  Triple-Duty  Emergency  Valve. 

383.  Pilot  Valve,  Anderson  Triple-Duty  Emergency  Valve. 

384.  Types  of  Check  Valves. 
385-7.   Types  of  Blow-off  Valves. 

388.  Blow-off   and   Feed-Water  Piping,   South   Side   Elevated  Railway  Station, 

Chicago,  Illinois. 

389.  Dead-Weight  Safety  Valve. 

390.  Common  Lever  Safety  Valve. 

391.  Consolidated  Pop  Safety  Valve. 

392.  Foster  Back-Pressure  Valve. 

393.  Davis  Back-Pressure  Valve. 

394.  Crane  Atmospheric  Relief  Valve. 

395.  Acton  Atmospheric  Relief  Valve. 


ILLUSTRATIONS  xxix 

FIG. 

396.  Kieley  Reducing  Valve. 

397.  Foster  Pressure  Regulator. 

398.  Types  of  Foot  Valves. 

CHAPTER  XVI.  —  LUBRICANTS  AND  LUBRICATION. 

399.  Oil  Cup  Lubrication. 

400.  Nugent's  Telescopic  Oiler. 

401.  Ring  Oiler. 

402.  Centrifugal  Oiler. 

403.  Pendulum  Oiler. 

404.  Simple  Gravity  Feed  System. 

405.  Low-Pressure  Gravity  Oil  Feed. 

406.  Compressed- Air  Oiling  System  at  First  National  Bank  Building,   Chicago, 

Illinois. 

407.  Leyland  Automatic  Oil  Cup. 

408.  Common  Sight-Feed  Hydrostatic  Lubricator. 

409.  Lunkenheimer  Sight-Feed  Lubricator. 

410.  Central  Hydrostatic  Lubricating  System. 

411.  Rochester  Forced-Feed  Lubricator,  Single  Feed. 

412.  Forced-Feed  Cylinder  Lubricator,  Multi-feed. 

413.  Siegrist  System. 

414.  Siegrist  Sight-Feed  Lubricator. 

415.  White  Star  Oil  Filter. 

416.  Turner  Oil  Filter. 

CHAPTER  XVII.  —  FINANCE  AND  ECONOMICS  —  COST  OF  POWER. 

417.  Influence  of  Load  Factor  on  Cost  of  Power. 

418.  Cost  of  Power,  Manufacturing  Plant. 

418  a,  b,  c.    Cost  of  Power  in  Large  Central  Stations. 

CHAPTER  XVIII.  —  TESTING  AND  MEASURING  INSTRUMENTS. 

419.  Piston  Water  Meter. 

420.  Disk  Water  Meter. 

421.  Venturi  Meter. 

422.  St.  John's  Steam  Meter. 
422a.   Burnham  Steam  Meter. 

423.  Different  Forms  of  Manometer  Draft  Gauges. 

424.  Bourdon  Pressure  Gauge. 

425.  Bristol  Recording  Air  Thermometer. 

426.  Bristol  Thermo-Electric  Pyrometer. 

427.  Element  for  Callendar  Resistance  Pyrometer. 

428.  Wanner  Optical  Pyrometer. 

429.  Fe"ry  Radiation  Pyrometer. 

430.  Orsat  Apparatus. 

431.  Arndt's  Econometer. 

432.  Ados  CO2  Recorder  —  Gas-Weighing  Apparatus. 

433.  Sarco  CO2  Recorder. 

434.  Separating  Calorimeter. 


ILLUSTRATIONS 
FIG. 

435.  Throttling  Calorimeter. 
435a.    Universal  Calorimeter. 

436.  Mahler  Bomb  Calorimeter. 

437.  Parr  Fuel  Calorimeter. 

CHAPTER  XX.  —  A  TYPICAL  STEAM  TURBINE  STATION. 

438.  General  Arrangement  of  Plant  and  Grounds. 

439.  North  Elevation  of  Building. 

440.  General  Plan  of  Boiler  and  Turbine  Room. 

441.  Section  through  Boiler  and  Turbine  Room. 

442.  Section  through  Boiler  Room. 

443.  General  Plan  Quarry  Street  Station. 

444.  Side  Elevation  of  Quarry  Street  Station. 

445.  Sectional  Elevation  of  Unit  No.  4  Quarry  Street  Station. 

CHAPTER  XXI.  —  A  TYPICAL  ISOLATED  STATION. 

446.  Plan  of  Ground  Floor. 

447.  Sectional  Elevation  through  Line  DD  of  Fig.  446. 

448.  Cross  Section  through  Line  CC  of  Fig.  446. 

449.  Longitudinal  Section  through  BB  of  Fig.  446. 

450.  Plan  of  Basement. 

451.  Diagram  of  Switchboard  Connections. 

APPENDIX  B.  —  A.  S.  M.  E.  RULES  FOR  CONDUCTING  BOILER  TESTS. 

452.  Ringlemann  Smoke  Chart. 

453.  Graphical  Log  Boiler  Test. 

APPENDIX  C.  —  A.  S.  M.  E.  RULES  FOR  CONDUCTING  STEAM  ENGINE  TESTS. 

454-5.   Rope  Brakes. 

456.  Alden  Absorption  Dynamometer. 

457.  Indicator  Card,  Simple  Engine. 

458.  Indicator  Card,  Four- Valve  Engine,  Slow  Speed. 

459.  Indicator  Card,  Single- Valve  Engine,  High  Speed. 
460-1.   Temperature-Entropy  Diagrams. 

APPENDIX  H.  —  MOLLIER  DIAGRAM. 

462.  General  Outline  of  Mollier's  Diagram. 

463.  Complete  Reproduction  of  Diagram. 


STEAM  POWER  PLANT  ENGINEERING 


CHAPTER   I. 

ELEMENTARY  STEAM  POWER  PLANTS. 

1.  General.  —  An  equipment  for  the  generation  of  power  is  known 
as  a  station  or  plant.  When  equipped  to  generate  electricity  for  the 
production  of  light  or  power  it  is  known  as  an  Electrical  Station  or 
Electric  Light  and  Power  Station.  The  term  Heating  Plant  refers  to 
a  plant  in  which  the  heat  energy  of  fuel  is  made  available  for  heating 
purposes  through  the  medium  of  steam  or  hot  water.  In  general, 
plants  or  stations  are  designated  according  to  the  manner  in  which  the 
energy  of  the  fuel  is  utilized. 

When  a  station  distributes  power  to  a  number  of  consumers  more 
or  less  distant,  it  is  called  a  Central  Station.  When  the  distances  are 
very  great,  electrical  current  of  high  tension  is  frequently  employed, 
and  is  transformed  and  distributed  at  convenient  points  through  Sub- 
stations. A  plant  designed  to  furnish  power  or  heat  to  a  building  or  a 
group  of  buildings  under  one  management  is  called  an  Isolated  Station. 
For  example,  the  power  plant  of  an  office  building  is  usually  called  an 
isolated  station. 

When  the  exhaust  steam  from  the  engines  is  discharged  at  approxi- 
mately atmospheric  pressure,  the  plant  is  said  to  be  operating  non- 
condensing.  When  the  exhaust  steam  is  condensed,  reducing  the  back 
pressure  on  the  piston  by  the  partial  vacuum  thus  formed,  the  plant 
is  said  to  operate  condensing. 

When  the  exhaust  steam  may  be  used  for  manufacturing,  heating,  or 
other  useful  purposes,  as  is  frequently  the  case  in  various  manufac- 
turing establishments,  and  in  large  office  buildings,  it  is  usually  more 
economical  to  run  non-condensing,  while  power  plants  for  electric 
lighting  and  power,  pumping  stations,  air  compressor  plants,  and  others, 
in  which  the  load  is  fairly  constant  and  the  exhaust  steam  is  not 
required  for  heating,  are  generally  operated  condensing. 


STE^M  POWER  PLANT  ENGINEERING 


r       B  ^--*-*-- /  Non- Condensing  Plant.  —  Fig.  1  gives  a  diagrammatic 

outline  of  the  essential  elements  of  the  simplest  form  of  steam  power 
plant.  The  equipment  is  complete  in  every  respect  and  embodies 
all  the  accessories  necessary  for  successful  operation.  Where  a  small 


FIG.  1.    Elementary  Non-Condensing  Plant. 

amount  of  power  is  desired  at  intermittent  periods,  as  in  hoisting 
systems,  threshing  outfits  and  traction  machinery,  the  arrangement 
is  substantially  as  illustrated.  The  output  in  these  cases  seldom 
exceeds  50  horse  power  and  the  time  of  operation  is  usually  short,  so 
the  cheapest  of  appliances  are  installed,  simplicity  and  low  first  cost 
being  more  important  than  economy  of  fuel. 

Such  a  plant  has  three  essential  elements:  (1)  The  furnace,  (2)  the 
boiler,  and  (3)  the  engine.  Fuel  is  fed  into  the  furnace,  where  it  is 
burned.  A  portion  of  the  heat  liberated  from  the  fuel  by  combustion 
is  absorbed  by  the  water  in  the  boiler,  converting  it  into  steam  under 
pressure.  The  steam  being  admitted  to  the  cylinder  of  the  engine 
does  work  upon  the  piston,  and  is  then  exhausted  through  a  suitable 
pipe  to  the  atmosphere.  The  process  is  a  continuous  one,  fuel  and 
water  being  fed  into  the  furnace  and  the  boiler  in  proportion  to  the 
power  demanded. 


ELEMENTARY  STEAM  POWER  PLANTS  3 

In  such  an  elementary  plant,  certain  accessories  are  necessary  for 
successful  operation.  The  grate  for  supporting  the  fuel  during  com- 
bustion consists  of  a  cast  iron  grid  or  of  a  number  of  cast  iron  bars 
spaced  in  such  a  manner  as  to  permit  the  passage  of  air  through  the 
fuel  from  below.  The  solid  waste  products  fall  through  or  are  "  sliced  " 
through  the  grate  bars  into  the  ash  pit,  from  which  they  may  be  removed 
through  the  ash  door.  The  latter  acts  also  as  a  means  of  regulating  the 
supply  of  air  below  the  grate.  Fuel  is  fed  into  the  furnace  through 
the  fire  door,  and  when  occasion  demands,  air  may  be  supplied  above  the 
bed  of  fuel  by  means  of  this  door.  The  combustion  chamber  is  the  sp»ace 
between  the  bed  of  fuel  and  the  boiler  heating  surface,  its  office  being  to 
afford  a  space  for  the  oxidation  of  the  combustible  gases  from  the  solid 
fuel  before  they  are  cooled  below  ignition  temperature  by  the  com- 
paratively cool  surfaces  of  the  boiler.  The  chimney  or  stack  discharges 
the  products  of  combustion  into  the  atmosphere,  and  serves  to  create 
the  draft  necessary  to  draw  the  air  through  the  bed  of  fuel.  Various 
forced  draft  appliances  are  sometimes  used  to  assist  or  to  entirely 
replace  the  chimney.  The  heating  surface  is  that  portion  of  the  boiler 
area  which  comes  into  contact  with  the  hot  furnace  gases,  absorbs  the 
heat  and  transmits  it  to  the  water.  In  the  small  plant,  illustrated  in 
Fig.  1,  the  major  portion  of  the  heating  surface  is  composed  of  a  number 
of  fire  tubes  below  the  water  line,  through  which  the  heated  gases 
pass.  The  volume  above  the  water  level  is  called  the  steam  space. 
Water  is  forced  into  the  boilers  either  by  a  feed  pump  or  an  injector. 
In  small  plants  of  the  type  considered,  steam  pumps  are  seldom  em- 
ployed; the  injector  answers  the  purpose  and  is  considerably  cheaper. 
A  safety  valve  connected  to  the  steam  space  of  the  boiler  automatically 
permits  steam  to  escape  to  the  atmosphere  if  an  excessive  pressure  is 
reached.  The  water  level  is  indicated  by  try  cocks  or  by  a  gauge  glass, 
the  top  of  which  is  connected  with  the  steam  space  and  the  bottom 
with  the  water  space.  Try  cocks  are  small  valves  placed  in  the  water 
column  or  boiler  shell,  one  at  normal  water  level,  one  above  it,  and  one 
below.  By  opening  the  valves  from  time  to  time  the  water  level  is 
approximately  ascertained.  They  are  ordinarily  used  in  case  of  acci- 
dent to  the  gauge  glass.  Fusible  plugs  are  frequently  inserted  in  the 
boiler  shell  at  the  lowest  permissible  water  level.  They  are  com- 
posed of  an  alloy  having  a  low  fusing  point  which  melts  when  in  con- 
tact with  steam,  thus  giving  warning  by  the  blast  of  the  escaping  steam 
if  the  water  level  gets  dangerously  low.  The  blowoff  cock  is  a  valve 
fitted  to  the  lowest  part  of  the  boiler  to  drain  it  of  water  or  to  discharge 
the  sediment  which  deposits  in  the  bottom.  The  steam  outlet  of  a 
boiler  is  usually  called  the  steam  nozzle. 


4  STEAM  POWER  PLANT  ENGINEERING 

The  essential  accessories  of  the  simple  steam  engine  include:  A 
throttle  valve  for  controlling  the  supply  of  steam  to  the  engine;  the 
governor,  which  regulates  the  speed  of  the  engine  by  governing  the  steam 
supply;  the  lubricator,  attached  to  the  steam  pipe,  which  is  usually  of 
the  "sight  feed"  class  and  provides  for  lubrication  of  piston  and 
valve.  Lubrication  of  the  various  bearings  is  effected  by  oil  cups 
suitably  located.  Drips  are  placed  wherever  a  water  pocket  is  apt  to 
form  in  order  that  the  condensation  may  be  drained.  The  apparatus 
to  be  driven  by  the  engine  may  be  direct  connected  to  the  crank  shaft 
or  belted  to  the  fly  wheel  or  geared. 

In  small  plants  of  this  type  no  attempt  is  made  to  utilize  the  exhaust 
steam  except  in  instances  where  the  stack  is  too  short  to  create  the 
necessary  draft,  in  which  case  the  exhaust  may  be  discharged  up  the 
stack.  If  the  draft  is  produced  by  convection  of  the  heated  gases  in 
the  chimney,  the  fuel  is  said  to  be  burned  under  natural  draft;  if  the 
natural  draft  is  assisted  by  the  exhaust  steam,  the  fuel  is  said  to  be 
burned  under  forced  draft.  The  power  realized  from  a  given  weight  of 
fuel  is  very  low  and  seldom  exceeds  2J  per  cent  of  the  heat  value  of  the 
fuel.  The  distribution  of  the  various  losses  in  a  plant  of,  say,  40  horse 
power  is  approximately  as  follows: 

B.T.U. 

Heat  value  of  1  pound  of  coal      14,500 

Boiler  and  furnace  losses,  50  per  cent 7,250 

Heat  of  the  steam,  50  per  cent . 7,250 

Heat  equivalent  of  one  horse  power  hour 2,545 

Heat  used  to  develop  one  horse  power  hour  (50  pounds  steam  per 
horse  power  hour,  pressure  80  pounds  gauge,  feed  water  62  degrees  F.)  57,500 

Per  cent. 

2  545 

Percentage  of  heat  in  the  steam,  realized  as  work,      '         ....          4.4 

o7,o(JO 

2  545 
Percentage  of  heat  value  of  the  coal  realized  as  work,  '  2.2 

O/,OUU  -r-  U.OU 


The  power  plant  of  the  modern  locomotive  is  very  much  like  that 
illustrated  in  Fig.  1,  the  main  difference  lying  in  the  type  of  boiler  and 
engine.  The  entire  exhaust  from  the  engine  is  discharged  up  the 
stack  through  a  suitable  nozzle,  since  the  extreme  rate  of  combustion 
requires  an  intense  draft.  The  engine  is  a  highly  efficient  one  compared 
to  that  in  the  illustration,  and  the  performance  of  the  boiler  is  more 
economical.  In  average  locomotive  practice  about  6  per  cent  of  the 
heat  value  of  the  fuel  is  converted  into  mechanical  energy  at  the  draw 
bar.  In  general,  a  non-condensing  steam  plant  in  which  the  heat  of 
the  exhaust  is  wasted  is  very  uneconomical  of  fuel,  even  under  the 


ELEMENTARY  STEAM  POWER  PLANTS  5 

most  favorable  conditions,  and  seldom  transforms  as  much  as  7  per  cent 
of  the  heat  value  of  the  fuel  into  mechanical  energy. 

3.  Non-Condensing  Plant.  Exhaust  Steam  Heating.  —  Fig.  2  gives 
a  diagrammatic  arrangement  of  a  simple  non-condensing  plant  differ- 
ing from  Fig.  1  in  that  the  exhaust  steam  is  used  for  heating  pur- 
poses. This  shows  the  essential  elements  and  accessories,  but  omits 
a  number  of  small  valves,  by-passes,  drains,  and  the  like  for  the  sake 
of  simplicity.  The  plant  is  assumed  to  be  of  sufficient  size  to  warrant 
the  installation  of  efficient  appliances.  Steam  is  led  from  the  boiler 
to  the  engine  by  the  steam  main.  The  moisture  is  removed  from  the 
steam  before  it  enters  the  cylinder  by  a  steam  separator.  The  moisture 
drained  from  the  separator  is  either  discharged  to  waste  or  returned  to 
the  boiler.  The  exhaust  steam  from  the  engine  is  discharged  into  the 
exhaust  main  where  it  mingles  with  the  steam  exhausted  from  the  steam 
pumps.  Since  the  exhaust  from  engines  and  pumps  contains  a  large 
portion  of  the  cylinder  oil  introduced  into  the  live  steam  for  lubricat- 
ing purposes,  it  passes  through  an  oil  separator  before  entering  the 
heating  system.  After  leaving  the  oil  separator  the  exhaust  steam 
is  diverted  into  two  paths,  part  of  it  entering  the  feed  water  heater  where 
it  condenses  and  gives  up  heat  to  the  feed  water,  and  the  remainder 
flowing  to  the  heating  system.  During  warm  weather  the  engine 
generally  exhausts  more  steam  than  is  necessary  for  heating  purposes, 
in  which  case  the  surplus  steam  is  automatically  discharged  to  the 
exhaust  head  through  the  back  pressure  valve.  The  back  pressure  valve 
is,  virtually,  a  large  weighted  check  valve  which  remains  closed  when 
the  pressure  in  the  heating  system  is  below  a  certain  prescribed  amount 
but  which  opens  automatically  when  the  pressure  is  greater  than  this 
amount.  During  cold  weather  it  often  happens  that  the  engine  exhaust 
is  insufficient  to  supply  the  heating  system,  the  radiators  condensing 
the  steam  more  rapidly  than  it  can  be  supplied.  In  this  case  live  steam 
from  the  boiler  is  automatically  fed  into  the  main  heating  supply  pipe 
through  the  reducing  valve. 

The  condensed  steam,  and  the  entrained  air  which  is  always  present, 
are  automatically  discharged  from  the  radiators  by  a  thermostatic  valve 
into  the  returns  header.  The  thermostatic  valve  is  so  constructed  that 
when  in  contact  with  the  comparatively  cool  water  of  condensation  it 
remains  open  and  when  in  contact  with  steam  it  closes.  The  vacuum 
pump  or  vapor  pump  exhausts  the  condensed  steam  and  air  from  the 
returns  header  and  discharges  them  to  the  returns  tank.  The  small 
pipe  S  admits  cold  water  to  the  vacuum  pump  and  serves  to  condense 
the  heated  vapor,  and  at  the  same  time  supply  the  necessary  make  up 
water  to  the  system.  The  returns  tank  is  open  to  the  atmosphere  so 


STEAM  POWER  PLANT  ENGINEERING 


ELEMENTARY  STEAM  POWER  PLANTS  7 

that  the  air  discharged  from  the  vacuum  pump  may  escape.  From 
the  returns  tank  the  condensed  steam  gravitates  to  the  feed  water 
heater  where  its  temperature  is  raised  to  practically  that  of  the  exhaust 
steam.  The  feed  water  gravitates  to  the  feed  pump  and  is  forced  into 
the  boiler.  There  are  several  systems  of  exhaust  steam  heating  in 
current  practice  which  differ  considerably  in  details,  but,  in  a  broad 
sense,  are  similar  to  the  one  just  described.  The  more  important  of 
these  will  be  described  later  on. 

During  the  summer  months  when  the  heating  system  is  shut  down, 
the  plant  operates  as  a  simple  non-condensing  station  and  practically 
all  of  the  exhaust  steam,  amounting  to  perhaps  60  per  cent  of  the  heat 
value  of  the  fuel,  is  wasted.  The  total  coal  consumption,  therefore,  is 
charged  against  the  power  developed.  During  the  winter  months, 
however,  all,  or  nearly  all  of  the  exhaust  steam  may  be  used  for  heating 
purposes  and  the  power  becomes  a  relatively  small  percentage  of  the 
total  fuel  energy  utilized.  The  percentage  of  heat  value  of  the  fuel 
chargeable  to  power  depends  upon  the  size  of  the  plant,  the  number 
and  character  of  engines  and  boilers,  and  the  conditions  of  operation. 
It  ranges  anywhere  from  50  per  cent  to  100  per  cent  for  the  summer 
months  and  may  run  as  low  as  6  per  cent  for  the  winter  months.  This  is 
on  the  assumption,  of  course,  that  the  engine  is  debited  only  with  the 
difference  between  the  coal  necessary  to  produce  the  heat  entering  the 
cylinder  and  that  utilized  in  the  heating  system. 

4.  Elementary  Condensing  Plant.  —  Under  the  most  favorable  con- 
ditions a  non-condensing  plant  can  never  be  expected  to  realize  more 
than  7  per  cent  of  the  heat  value  of  the  fuel  as  power.  In  large  non- 
condensing  power  stations  the  demand  for  exhaust  steam  is  usually 
limited  to  the  heating  of  the  feed  water,  and  as  only  12  per  cent  or 
15  per  cent  can  be  utilized  in  this  manner,  the  greater  portion  of  the 
heat  in  the  exhaust  is  lost.  Non-condensing  engines  require  from  20  to 
60  pounds  of  steam  per  hour  for  each  horse  power  developed.  On  the 
other  hand  in  condensing  engines  the  steam  consumption  may  be  reduced 
to  as  low  as  10  pounds  per  horse  power  hour.  The  saving  of  fuel 
is  at  once  apparent. 

Fig.  3  gives  a  diagrammatic  arrangement  of  a  simple  condensing 
plant  in  which  the  back  pressure  on  the  engine  is  reduced  by  condens- 
ing the  exhaust  steam.  A  different  type  of  boiler  from  that  in  Fig.  1 
or  Fig.  2  has  been  selected  for  the  purpose  of  bringing  out  a  few  of 
the  characteristic  elements.  The  products  of  combustion  instead  of 
passing  directly  through  fire  tubes  to  the  stack  as  in  Fig.  1  are  deflected 
back  and  forth  across  a  number  of  water  tubes,  by  the  bridge  wall  and  a 
series  of  baffles.  After  imparting  the  greater  part  of  their  heat  to  the 


STEAM  POWER  PLANT  ENGINEERING 


ELEMENTARY  STEAM  POWER  PLANTS  9 

heating  surface  the  products  of  combustion  escape  to  the  chimney 
through  the  breeching  or  flue.  The  rate  of  flow  is  regulated  by  a  damper 
placed  in  the  breeching  as  indicated. 

The  steam  generated  in  the  boiler  is  led  to  the  engine  through  the 
main  header.  The  steam  is  exhausted  into  a  condenser  in  which  its 
latent  heat  is  absorbed  by  injection  or  cooling  water.  The  process 
condenses  the  steam  and  creates  a  partial  vacuum.  The  condensed 
steam,  injection  water,  and  the  air  which  is  invariably  present  are 
withdrawn  by  an  air  pump  and  discharged  to  the  hot  well.  In  case  the 
vacuum  should  fail  as  by  stoppage  of  the  air  pump  the  exhaust  steam 
is  automatically  discharged  to  the  exhaust  head  by  the  atmospheric 
relief  valve,  and  the  engine  will  operate  non-condensing.  The  atmos- 
pheric relief  valve  is  a  large  check  valve  which  is  held  closed  by  atmos- 
pheric pressure  as  long  as  there  is  a  vacuum  in  the  condenser.  When 
the  vacuum  fails  the  pressure  of  the  exhaust  becomes  greater  than 
that  of  the  atmosphere  and  the  valve  opens. 

The  feed  water  may  be  taken  from  the  hot  well  or  from  any  other 
source  of  supply  and  forced  into  the  heater.  In  this  particular  case  it  is 
taken  from  a  cold  supply  and  upon  entering  the  heater  is  heated  by  the 
exhaust  steam  from  the  air  and  feed  pumps.  From  the  heater  it 
gravitates  to  the  feed  pump  and  is  forced  into  the  boiler.  Various 
other  combinations  of  heaters,  pumps,  and  condensers  are  necessary 
in  many  cases,  depending  upon  the  conditions  of  operation.  Feed 
pumps,  air  pumps,  and  in  fact  all  small  engines  used  in  connection  with 
a  steam  power  plant  are  usually  called  auxiliaries. 

A  well-designed  station  similar  to  the  one  illustrated  in  Fig.  3  is 
capable  of  converting  about  10  per  cent  of  the  heat  value  of  the  fuel 
into  mechanical  energy.  The  various  heat  losses  are  approximately 
as  follows: 

BOILER  LOSSES.  Per  Cent. 

Loss  due  to  fuel  falling  through  the  grate      . 2 

Loss  due  to  incomplete  combustion 2 

Loss  due  to  heat  carried  away  in  chimney  gases 23 

Radiation  and  other  losses 8 

Total ';.     35 

Heat  used  by  engines  and  auxiliaries  (16  pounds  of  steam  per 

I.H.P.  hour,  pressure  150  pounds,  feed  water  210°  F.)   .    .    .    .  16,250 

Engine  and  generator  friction,  5  per  cent 812 

Leakage,  radiation,  etc.,  2  per  cent 325 

Total 17,387 

Heat  equivalent  of  one  electrical  horse  power 2,545 

Percentage  of  the  heat  value  of  the  steam  converted  into  electrical       Per  Cent* 
energy       14 .7 

Percentage  of  heat  value  of  fuel  converted  into  electrical  energy 
2545  X  0.65 

17,387  y'5 


10  STEAM  POWER  PLANT  ENGINEERING 

5.  Condensing  Plant  with  Full  Complement  of  Heat-saving  Appli- 
ances. —  When  fuel  is  costly  it  frequently  becomes  necessary  for  the 
sake  of  economy  to  reduce  the  heat  wastes  as  much  as  possible.  The 
chimney  gases,  which  in  average  practice  are  discharged  at  a  tem- 
perature between  450  degrees  F.  and  550  degrees  F.,  represent  a  loss  of 
20  per  cent  to  30  per  cent  of  the  total  value  of  the  fuel.  If  part  of  the 
heat  could  be  reclaimed  without  impairing  the  draft  the  gain  would 
be  directly  proportional  to  the  reduction  in  temperature  of  the  gases. 
Again,  in  some  types  of  condensers  all  of  the  steam  exhausted  by  the 
engine  is  condensed  by  the  circulating  water  and  discharged  to  waste. 
If  provision  could  be  made  for  utilizing  part  of  the  exhaust  steam  for 
feed-water  heating,  the  efficiency  of  the  plant  could  be  correspondingly 
increased.  In  many  cases  the  cost  of  installing  such  heat-saving  devices 
would  more  than  offset  the  gain  effected,  but  occasions  arise  where  they 
give  marked  economy. 

Fig.  4  gives  a  diagrammatic  arrangement  of  a  condensing  plant  in 
which  a  number  of  heat-reclaiming  devices  are  installed.  The  plant  is 
assumed  to  consist  of  a  number  of  engines,  boilers,  and  auxiliaries. 
Coal  is  automatically  transferred  from  the  cars  to  coal  hoppers  placed 
above  the  boiler,  by  a  system  of  buckets  and  conveyors.  These  hoppers 
store  the  coal  in  sufficient  quantities  to  keep  the  boiler  in  continu- 
ous operation  for  some  time.  From  the  hoppers  the  coal  is  fed 
intermittently  to  the  stoker  by  means  of  a  down  spout.  The  stoker 
feeds  the  furnace  in  proportion  to  the  power  demanded  and  auto- 
matically rejects  the  ash  and  refuse  to  the  ash  pit.  The  ashes  are 
removed  from  the  ash  pit  when  occasion  demands,  and  are  transferred 
to  the  ash  hopper  by  the  same  system  of  buckets  and  conveyor  which 
handles  the  coal.  The  ash  hopper  is  usually  placed  alongside  the  coal 
hoppers  and  is  not  unlike  them  in  general  appearance  and  construction. 

The  products  of  combustion  are  discharged  to  the  stack  through  the 
flue  or  breeching.  Within  the  flue  is  placed  a  feed-water  heater  called 
an  economizer,  the  function  of  which  is  to  absorb  part  of  the  heat  from 
the  gases  on  their  way  to  the  chimney.  The  heat  reclaimed  by  the 
economizer  varies  widely  with  the  conditions  of  operation  and  ranges 
between  5  per  cent  and  20  per  cent.  Since  the  economizer  acts  as  a 
resistance  to  the  passage  of  the  products  of  combustion  it  is  sometimes 
necessary  to  increase  the  draft  either  by  increasing  the  height  of  the 
chimney  or,  as  is  the  usual  practice,  by  using  a  forced  draft  system. 

Part  of  the  heat  of  the  exhaust  steam  is  reclaimed  by  a  vacuum  heater 
which  is  placed  in  the  exhaust  line  between  engine  and  condenser. 
For  example,  if  the  feed  water  has  a  normal  temperature  of  60  degrees  F. 
and  the  vacuum  in  the  condenser  is  26  inches,  the  vacuum  heater  will 


ELEMENTARY  STEAM  POWER  PLANTS 


11 


12          STEAM  POWER  PLANT  ENGINEERING 

raise  the  temperature  of  the  feed  to  say  120  degrees  F.,  thereby 
effecting  a  gain  in  heat  of  approximately  6  per  cent.  If  the  feed  supply 
is  taken  from  the  hot  well  the  vacuum  heater  is  without  purpose,  as  the 
temperature  of  the  hot  well  will  not  be  far  from  120  degrees  F. 

Referring  to  the  diagram,  the  path  of  the  steam  is  as  follows:  From 
the  boiler  it  flows  through  the  boiler  lead  to  the  main  header  or  equalizing 
pipe.  From  the  main  header  it  flows  through  the  engine  lead  to  the 
high-pressure  cylinder.  The  exhaust  steam  discharges  from  the  low- 
pressure  cylinder  through  the  vacuum  heater  and  into  the  condenser. 
Part  of  the  exhaust  steam  is  condensed  in  the  vacuum  heater  and  gives 
up  its  latent  heat  to  the  feed  water.  The  remainder  is  condensed  by 
the  injection  water  which  is  forced  into  the  condenser  chamber  by  the 
circulating  pump.  The  condensed  steam  and  circulating  water  gravitate 
through  the  tail  pipe  to  the  hot  well.  The  air  which  enters  the  con- 
denser either  as  leakage  or  entrainment  is  withdrawn  by  the  air  pump. 
The  steam  exhausted  by  the  feed  pump,  air  pump,  stoker  engine,  and 
other  steam-driven  auxiliaries  is  usually  discharged  into  the  atmospheric 
heater,  which  still  further  heats  the  feed  water. 

Referring  to  the  feed  water,  the  circuit  is  as  follows:  The  pump 
draws  in  cold  water  at  a  temperature  of  say  60  degrees  F.,  and  forces 
it  in  turn  through  the  vacuum  heater,  the  atmospheric  heater,  and  the 
economizer  into  the  boiler.  The  vacuum  heater  raises  the  temperature 
to  120  degrees  F.,  the  atmospheric  heater  increases  it  to  212  degrees  F., 
and  the  economizer  still  further  to  about  300  degrees  F.  The  heat 
reclaimed  by  this  series  of  heaters  is  evidently  the  equivalent  of  that 
necessary  to  raise  the  feed  water  from  60  degrees  F.  to  300  degrees  F., 
or  approximately  24  per  cent  of  the  total  heat  supplied.  In  some 
plants  the  economizer  only  is  installed,  in  others  the  economizer  and 
atmospheric  heater  are  deemed  desirable,  still  others  utilize  all  three. 
The  distribution  of  the  heat  losses  in  a  plant  of  this  type  operating  under 
favorable  conditions  is  approximately  as  follows: 


Per  Cent.  B.T.U. 
Delivered  to  engine,  15  pounds  steam  per  I.H.P.  hour; 

pressure  150  pounds,  feed  60°  F 100  17,482 

Delivered  to  feed  pump 1.5  262 

Delivered  to  circulating  pump 1.5  262 

Delivered  to  air  pump 

Delivered  to  small  auxiliaries 1.5  262 

Loss  in  leakage  and  drips 0.5  87 

Engine  and  generator  friction 5 

Radiation  and  minor  losses 1  175 

Total  .                                                          1V53 


ELEMENTARY  STEAM  POWER  PLANTS         13 

PerCent.  B.T.U. 

Returned  by  vacuum  heater      5.5  1,086 

Returned  by  atmospheric  heater .    .         7.9  1,560 

Returned  by  economizer 9.7  1,916 

Total 23.1        4,562 

Net  heat  delivered  to  engine  in  the  form  of  steam  to  pro- 
duce one  electrical  horse  power,  19,753-4,562  ....  15,191 

2545 
Percentage  converted  to  electrical  power  ....       16 . 7 

I  ' ),  1-M. 

Boiler  efficiency 70 

Percentage  of  heat  value  of  fuel  necessary  to  produce  one 

. .  ,  ,         ,  2545  X  0.70  , ,   _ 

electrical  horse  power  at  switchboard .    .       11.7 

lo,  1J1 

The  preceding  figures  give  the  results  of  very  good  practice.  So 
much  depends  upon  the  size  and  character  of  the  prime  movers,  the 
nature  of  the  fuel,  and  the  conditions  of  operation  that  no  definite 
figure  can  be  given  for  the  percentage  of  heat  converted  to  power  in  a 
given  type  of  station.  Six  per  cent  represents  good  average  practice 
in  a  non-condensing  plant  and  10  per  cent  in  a  condensing  plant. 
Pumping  stations  operating  continuously  under  full  load  have  realized 
as  much  as  15  per  cent  of  the  total  heat  value  of  the  fuel,  but  such 
performances  are  practically  unobtainable  in  connection  with  steam- 
driven  electrical  power  plants.  Steam  power  plants  as  a  class  are  very 
wasteful  of  fuel  at  the  best. 

One  of  the  best  recorded  performances  to  date  (March,  1909)  of  a 
steam-electric  power  plant  is  that  of  the  Pacific  Light  and  Power 
Company  at  Redondo,  Cal.  When  operating  under  regular  commer- 
cial conditions  approximately  14  per  cent  of  the  available  heat  of  the 
fuel  (crude  oil)  is  realized  as  power  at  the  switchboard.  For  a  detailed 
description  of  the  plant  and  the  results  of  the  acceptance  tests,  see 
Jour,  of  Elec.  Gas  and  Power,  Aug.  22,  1908. 


CHAPTER  IT. 

FUELS  AND   COMBUSTION. 

6.  General.  —  The  subject  of  fuels   and   combustion  has  been  so 
extensively  treated  by  various  authorities  that  a  comprehensive  dis- 
cussion would  be  without  purpose  here,  but  in  order  to  bring  out  more 
clearly  the  matter  pertaining  to  the  commercial  design  and  operation 
of  steam  power  plants  a  few  of  the  essential  elements  will  be  briefly 
treated. 

The  fuels  used  for  steam  making  are  coal,  coke,  wood,  peat,  mineral 
oil,  natural  and  artificial  gases,  refuse  products  such  as  straw,  manure, 
sawdust,  tan  bark,  bagasse,  and  occasionally  corn  and  molasses. 

In  most  cases  that  fuel  is  selected  which  develops  the  required  power 
at  the  lowest  cost,  taking  into  consideration  all  of  the  circumstances 
that  may  affect  its  use.  Occasionally  the  disposition  of  waste  products 
is  a  factor  in  the  choice,  but  such  instances  are  uncommon.  The  boilers 
and  furnaces  are  designed  to  suit  the  fuel  selected. 

7.  Classification  of  Fuels.  —  Fuels  may  be  divided  into  three  classes 
as  follows: 

1.  Solid  fuels. 

a.  Natural  fuels:  straw,  wood,  peat,  coal. 

6.  Prepared:  charcoal,  coke,  peat  and  other  briquettes. 

2.  Liquid  fuels. 

a.  Natural:  crude  oils. 

b.  Prepared:  distilled  oils,  alcohol,  molasses. 

3.  Gaseous  fuels. 

a.  Natural:  natural  gas. 

6.  Prepared:  coal  gas,  water  gas,  producer  gas,  oil  gas. 

8.  Solid  Fuels.  —  Solid  fuels  are  of  vegetable  origin  and  exist  in  a 
variety  of  forms  between  that  of  a  comparatively  recent  cellulose  growth 
and  that  of  nearly  pure  carbon  as  anthracite  coal.    They  owe  their  forms 
to  the  conditions  under  which  they  were  created  or  to  the  geological 
changes  which  they  have  undergone.     With  each  succeeding  stage  the 
percentage  of  carbon  increases.      The  chemical  changes  are  approxi- 
mately as  follows: 

14 


FUELS  AND  COMBUSTION 


15 


Substance. 

Carbon. 

Hydrogen. 

Oxygen. 

Pure  cellulose    

Per  Cent 
44.44 

Per  Cent 
6  17 

Per  Cent 
49  39 

Wood    

52.65 

5.25 

42  10 

Peat  

59.57 

5.96 

34  47 

Lignite     

66.04 

5.27 

28.69 

Brown  coal 

73  18 

5  58 

21  14 

Bituminous  coal                                                         .    . 

75  06 

5  84 

19  10 

Semi-bituminous  coal. 

89  29 

5  05 

6  66 

Anthracite      .    .           .        .        .    .        .... 

91  58 

3  96 

4  46 

Graphite      ... 

100.00 

All  natural  solid  fuels  contain  more  or  less  earthy  or  inorganic  matter 
which  is  not  combustible  and  therefore  remains  as  ash,  while  the 
organic  matter  is  consumed.  Sometimes  the  percentage  of  ash  is  so 
great  as  to  render  them  valueless  for  steam-making  purposes. 

Origin  and  Formation  of  Fuel:  Engng,  Aug.  23,  1901;  Am.  Geol.,  Feb.,  1899;  Col. 
Guard,  Sept.  10, 1897,  Oct.  1,  1897,  Jan.  14,  1898,  Jan.  28,  1898,  March  18,  1898,  Sept. 
14,  1900;  EC.  Geol.,  Oct.,  1905;  Eng.  U.S.,  April  1,  1903;  Ir.  and  Coal  Td.  Review, 
Feb.  4,  1898,  July  13,  1906. 

9.  Coal. — Coals  are  most  satisfactorily  classified  according  to  the 
constituents  of  the  combustible,  as 


Fixed  Carbon. 

Volatile  Matter. 

Anthracite  

Per  Cent 
97      to  92  5 

Per  Cent 
3      to    7  5 

Semi-anthracite    

92  5  to  87  5 

7  5  to  12  5 

Semi-bituminous  .    .    .•  

87.5  to  75 

12  5  to  25 

Bituminous,  Eastern  

75      to  60 

25      to  40 

Bituminous  Western 

65      to  50 

35      to  50 

Lignite 

Under  50 

Over  50 

Classification  of  Coals:  Am.  Inst.  of  Min.  Engrs.,  May,  1906,  Sept.,  1905;  Mines  and 
Minerals,  Dec.,  1906;  Min.  Rept.,  Apr.  26,  1906;  Col.  Guard,  July  6,  1900;  Power, 
Oct.,  1906. 

10.  Anthracite.  —  This  is  the  most  perfect  form  of  coal  and  consists 
almost  entirely  of  carbon;  it  contains  very  little  hydrocarbon  and  burns 
with  little  or  no  smoke,  is  slow  to  ignite,  burns  slowly,  and  breaks  into 
small  pieces  when  rapidly  heated.  It  requires  a  very  large  grate  of 
about  twice  the  surface  necessary  for  bituminous  coal.  Large  sizes 
may  be  burned  in  almost  any  kind  of  a  furnace  and  with  moderate  draft. 
For  small  sizes  a  thinner  bed  has  to  be  carried  unless  a  strong  draft  is 
used.  There  is  difficulty  in  keeping  it  free  from  air  holes.  When 


16  STEAM  POWER  PLANT   ENGINEERING 

possible,  the  coal  should  be  at  least  six  inches  deep  on  the  grates.  On 
account  of  the  large  percentage  of  ash  in  the  smaller  size,  the  fire 
requires  frequent  cleaning.  Anthracite  does  not  require  "  slicing " 
and  should  be  disturbed  only  when  cleaning  is  necessary. 

Small  Size  Anthracite :  Eng.  and  Min.  Jour.,  Dec.  22,  1904.  Heat  Value  of  Anthra- 
cite, Small  Sizes,  and  the  Best  Way  of  burning  it :  Col.  Guard,  Nov.  26,  1897.  Anthra- 
cite Coal  Mines  and  Coal  Mining :  Rev.  of  Rev.,  July,  1902.  The  Screening  of  Anthra- 
cite :  Col.  Guard,  Sept.  20,  1901.  Preparation  of  Anthracite  :  Mines  and  Min.,  March, 
1905.  Anthracite  Washeries :  Am.  Inst.  of  Min.  Engrs.,  Nov.,  1905.  Anthracite 
Coal  Fields  of  Pennsylvania:  Min.  Mag.,  March,  1905.  Virginia  Anthracite :  Eng. 
News,  Oct.  20,  1904;  Mines  and  Min.,  March,  1906.  Burning  of  Anthracite  Culm  of 
Poor  Quality  :  Trans.  A.S.M.E.,  7-390.  The  Use  of  Electricity  in  Anthracite  Mining  : 
Eng.  and  Min.  Jour.,  Feb.  2,  1907. 

Anthracite  is  classed  and  marketed  according  to  sizes,  the  following 
division  of  mesh  being  adopted  as  standard  at  Wilkesbarre  in  1891 : 

Egg  coal  must  pass  through  2|  inch  mesh  and  not  through   2  inch 

Stove  "         "      "  "2       "         "         "      "         "         1£  " 

Chestnut          "         "      "  "       \\     "         "         "      "         "  J  " 

pea  «         (i      u  u         a     «         «         it      n         u  i  K 

Buckwheat      "         "      "  "         %     "         "         "      "  £  " 

Rice  "         "      "        >*'><:      i     "         "         "      "         "  i  " 

Sizes  over  "  pea  coal."  are  prohibitive  in  price  for  steam  power  plant 
use  and  consequently  the  demand  is  limited  to  the  smaller  sizes. 

11.  Semi-Anthracite.  —  This  coal  kindles   more   readily    and   burns 
more  rapidly  than  anthracite.     It  requires  little  attention,  burns  freely 
with  a  short  flame,  and  yields  great  heat  with  little  clinker  and  ash.     It 
is  apt  to  split  up  on  burning  and  waste  somewhat  in  falling  through  the 
grates.     It  swells  considerably  but  does  not  cake.     It  has  less  density, 
hardness,  and  metallic  luster  than  anthracite,  and  can  generally  be  dis- 
tinguished by  its  tendency  to  soil  the  hands  while  pure  anthracite  will 
not. 

12.  Semi-Bituminous.  —  This    coal    is    softer  than  semi-anthracite. 
It  ignites  easily  and  burns  freely  under  a  moderate  draft.     It  gives  an 
intense  fire  and  is  an  excellent  steam  coal,  but  is  apt  to  smoke  con- 
siderably unless  special  provision  is  made  to  prevent  it. 

13.  Bituminous.  —  This  coal  contains  a  large  and  varying  amount  of 
volatile  matter  and  requires  careful  firing  to  prevent  smoke  and  clinker. 
Its  physical  properties  vary  widely,  so  much  so  that  it  is  usually  divided 
into  three  grades: 

1.  Dry  Bituminous  coal  is  sometimes  known  as  the  free-burning 
bituminous.  It  is  hard  and  dense,  black  in  color,  but  brittle  and 


FUELS  AND  COMBUSTION  17 

splintery.     It  ignites  somewhat  slowly,  burns  freely  with  a  short,  clean, 
bluish  flame,  little  smoke,  and  without  caking. 

2.  Bituminous  Caking  coal  swells  up,  becomes  pasty,  and  fuses  together 
in  burning.     It  contains  less  fixed  carbon  and  more  volatile  matter  than 
the  free-burning  grades.     Caking  coal  is  rich  in  hydrocarbon    and   is 
particularly    adapted  to  gas    making.     The    flame    is  of    a   yellowish 
color. 

3.  Long  Flaming  Bituminous  coal  is  similar  in  many  respects  to  the 
caking  coal  but  contains  a  larger  percentage  of  volatile  matter.     It  is 
free  burning,  with  a  long,  yellowish  flame.     It  may  be  either  caking  or 
non-caking. 

COAL    FIELDS    OF    THE    UNITED    STATES. 

•Alabama:  Mines  and  Minerals,  May,  1901. 

Arkansas :  Eng.  and  Min.  Jour.,  Sept.  12,  1903,  Oct.  28,  1905. 

Colorado:  Min.  Rept.,  Jan.  19,  1905,  March  2,  1905;  Jour.  W.  S.  Engrs.,  Dec., 
1903;  Mines  and  Minerals,  May,  1905. 

Illinois:  Min.  Mag.,  March,  1905;  Eng.  and  Min.  Jour.,  Jan.  13,  1906. 

Indiana:  Eng.  Rec.,  Jan.  27,  1906;  Min.  Mag.,  March,  1905;  Power,  July  and 
Aug.,  1902. 

Indian  Territory  :  Min.  Rept.,  May  17,  1906. 

Kansas:  Eng.  and  Min.  Jour.,  Dec.  7,  1901. 

Kentucky :  Col.  Guard,  Sept.  7,  1900. 

Michigan:  Eng.  and  Min.  Jour.,  June  30,  1900;  Min.  World,  Feb.  9,  1907. 

Missouri :  Am.  Inst.  of  Min.  Engrs.,  Jan.,  1905. 

Montana:  Min.  Mag.,  March,  1905;  Min.  World,  Nov.  24,  1906. 

Ohio :  Min.  Mag.,  March,  1905. 

Pennsylvania:  Eng.  and  Min.  Jour.,  Aug.  24,  1901;  Trans.  A.S.M.E.,  4-217;  Min. 
Mag.,  March,  1905;  Pro.  Eng.  S.  W.  Penn.,  Jan.,  1907. 

Texas :  Mines  and  Min.,  Oct.,  1905. 

Virginia:  Mines  and  Min.,  March,  1906;  Eng.  News,  Oct.  20,  1904. 

West  Virginia :  Eng.  and  Min.  Jour.,  May  12,  1904. 

Wyoming  :  Min.  World,  May  6,  1905. 


GENERAL. 

Coal  Mines  of  the  United  States :  Peabody  Atlas,  A.  Bement,  Chicago,  111.,  Min. 
World,  May  6,  1905. 

Coal  Resources  of  the  Pacific :  Eng.  Mag.,  May,  1902. 

Rocky  Mountain  Coal  Fields:  Min.  Rept.,  Jan.  5,  1905;  Jour.  Asso.  Eng.  Soc., 
Dec.,  1902. 

Coal  Fields,  U.S.  Northwest :  Rev.  of  Rev.,  Feb.,  1903. 

Coal  Fields,  U.S.  Southwest :  Eng.  and  Min.  Jour.,  Oct.  17,  1903. 

U.S.  Coal  Fields :  Steam  Boiler  Practice,  Wm.  Kent. 

Report  of  Coal  Testing  Plant :  U.S.  Geological  Survey,  Washington,  D.C. 


18  STEAM   POWER  PLANT  ENGINEERING 

14.  Lignite,  or  brown  coal,  is  a  substance  of  more  recent  geological 
formation  than  coal  and  represents  a  stage  in  development  intermediate 
between  coal  and  peat.     Its  specific  gravity  is  low,  1.2,  and  when  freshly 
mined  it  contains  as  high  as  50  per  cent  of  moisture.     It  is  non-caking 
and  gives  a  bright  but  slightly  smoky  flame.     It  is  a  low-grade  fuel  and 
is  used  where  good  coal  is  difficult  to  get.     Exposure  to  the  air  causes 
it  to  split  into  fine  pieces  like  air-slacked  lime.     It  is  very  fragile  and 
will  not  bear  much  handling  in  transportation. 

Eng.  and  Min.  Jour.,  Nov.  22,  1902,  Feb.  7,  1903;  Mines  and  Min.,  July,  1901. 
Lignite  of  Northeastern  Wyoming  :  Mines  and  Min.,  Feb.,  1907. 

15.  Peat,  or  Turf,  is  formed  by  the  slow  carbonization  under  water 
of  a  variety  of  accumulated  vegetable  materials.     It  is  unsuitable  for 
fuel  until  dried.     Peat  as  ordinarily  cut  and  dried  is  too  bulky  for  com- 
mercial competition  with  coal,  and  is  used  only  where  coal  is  prohibitive 
in  price.     When  properly  prepared  and  compressed  into  briquettes  peat 
is  an  excellent  fuel.     In  Russia,  Germany,  and  Holland  peat  briquettes 
have  passed  the   experimental  stage  and  several    millions  of  pounds 
are  manufactured  annually.     Peat  is  used  but  little  in  this  country 
at  present,  but  its  possibilities  are  beginning  to  attract  the  attention 
of  engineers.     The  proportion  in  which  the  various  primary  constit- 
uents exist  in  dried  peat  is  approximately  as  follows: 

Per  Cent. 

Fixed  carbon 35 

Volatile  matter 60 

Ash 5 

Peat:  Power,  Sept.,  1907;  Engr.  U.S.,  April  1,  1905;  Min.  World,  Sept.  30,  1905; 
Col.  Guard,  Nov.  30,  1900;  Mines  and  Min.,  July,  1901 ;  Eng.  and  Min.  Jour.,  Nov.  22, 
1902;  Feb.  7,  1903,  Eng.  Rec.,  52-191;  Sci.  Am.  Sup.,  March  2,  1907;  Elec.  Engr., 
Lond.,  Dec.  6,  1907. 

Fuel  Briquetting :  Jour.  Assn.  Eng.  Soc.,  Jan.,  1906;  Iron  Age,  April  19,  1906; 
Power,  Dec.,  1902;  Eng.  and  Min.  Jour.,  Nov.  8,  1902;  Mines  and  Min.,  Oct.,  1904; 
Power,  March,  1905;  Engr.  U.S.,  May,  1905. 

16.  Wood,  Straw,  Sawdust,  Bagasse,  Tanbark.  —  In   certain   locali- 
ties cord  wood  is  still  used  as  a  fuel,  but  the  steadily  increasing  values 
of  even  the  poorest  qualities  are  rapidly  prohibiting  its  use  for  steam- 
generating    purposes.     Sawdust,  shavings,  tanbark    and    other  waste 
products  of  wood  are  burned  under  boilers  in  situations  where  such 
disposition  nets  the  best  financial  returns.     Recent  progress,  however, 
in  industrial  chemistry  shows  that  ethyl  and  wood  alcohols  and  other 
valuable  by-products  can  be  cheaply  made  from  sawdust,  shavings, 


FUELS  AND  COMBUSTION 


19 


slashings  and  similar  waste  material,  and  it  is  not  unlikely  that  their 
use  for  steaming  purposes  will  be  unheard  of  in  a  comparatively  few 
years.  Table  00  gives  the  physical  and  chemical  characteristics  of  a 
number  of  woods. 

Wood  as  Fuel:  Prac.  Engr.  U.  S.,  Jan.,  1910,  p.  805;  Power  &  Engr.,  June  30,  1908, 
p.  1015;  Power,  Dec.,  1908,  p.  772. 

Burning  Sawdust:  Prac.  Engr.  U.  S.,  Jan..  1910,  p.  48;  Power  &  Engr.,  April  7, 
1908,  p.  536;  Oct.  13,  1908,  p.  613;  Jour,  of  Elec.,  Oct.,  1905. 


TABLE  00. 

PHYSICAL  AND  CHEMICAL  PROPERTIES  OF  WOODS,   STRAW  AND  TANBARK. 
(Prac.  Engr.  U.  S.,  Jan.,  1910.) 


Si 

•§  ri 

0  -o 

&! 

Il 
*& 

•d 

&  '• 
|J 

£  PH 
be 

1 

Equivalent  Weight 
of  Coal.  13,500 
B.T.U. 

1 

J« 

Hydrogen. 
Per  Cent. 

J 
1* 

B 

J 

tuO    &5 

% 

.* 
| 

s  1  • 

•3*-g 

0*    § 

«nfc 
gffl 

>> 

a 

< 

Ash 

46 

3520 

1420 

5450 

Hutton 

Beech    
Birch  

43 
45 
42 

3250 
2880 
3140 

1300 
1190 
1260 

49.36 
50.20 

6.01 
6.20 

42.69 
41.62 

0.91 
1.15 

1.06 
0.81 

5400 
5580 
5420 

Sharpless 
Hutton 

Chestnut 

41 

2350 

940 

5400 

E'm 

35 

2350 

940 

5400 

<.' 

Hemlock 

25 

1220 

580 

6410 

Hutton 

Hickory 

53 

4500 

1800 

5400 

Maple   Hard 

49 

3310 

1340 

5460 

Hutton 

Oak   Live 

59 

3850 

1560 

5460 

n 

"    White. 
"     Red 

52 
45 

3850 
3310 

1540 
1340 

49.64 

5.92 

41.16 

1.29 

1.97 

5400 
5460 

Rankine 
Hutton 

Pine  White 

25 

1920 

970 

6830 

it 

"    Yellow 

36 

2130 

1050 

6660 

ft 

Poplar  
Spruce 

36 
25 

2130 
1920 

1050 
970 

49.37 

6.21 

41.60 

0.96 

1.86 

6660 
6830 

ii 
K 

Walnut 

35 

3310 

1340 

5460 

t( 

Willow    .... 

25 

1920 

970 

49.96 

5.96 

39.56 

0.96 

3.37 

6830 

Rankine 

Average.  . 

49.70 

6.06 

41.30 

1.05 

1.80 

Straw  : 
Wheat  .  .  . 

* 

00 

Water 
16.00 

35.86 

5.01 

37.68 

0.45 

5.00 

Clark 

Barley  .  .  . 

o 

+9 

15.50 

36.27 

5.07 

38.26 

0.40 

4.50 

« 

Average 

to 

15.75 

36.06 

5.04 

37.97 

0.42 

4.75 

5155 

Tanbark 
Dry 

51.80 

6.04 

40.74 

1.42 

9500 

Myers 

Compressed. 


20 


STEAM  POWER  PLANT  ENGINEERING 


Bagasse,  or  megass,  is  refuse  sugar  cane  and  is  used  as  a  fuel  on  the 
sugar  plantations.  Its  heat  value  depends  upon  the  proportions  of 
fiber,  molasses,  sugar  and  water  left  after  the  extraction.  The  heat 
furnished  by  the  different  constituents  is  about  as  follows:  Fiber, 
8325  B.T.U.  per  pound;  sugar,  7223  B.T.U.  per  pound;  and  molasses, 
6956  B.T.U.  per  pound.  Table  0  gives  the  heat  value  of  bagasse  and 


TABLE  0. 

HEAT   VALUES  OF  BAGASSE  AND   VARIATION  WITH    DEGREE  OF    EXTRACTION. 


I  S 

0 

, 

2  •*•* 

lla 

i 

EJ 

-g    cS 

3 

Fiber. 

Sugar. 

Molasses. 

ID 

a; 

c.    . 

*->     0) 

£ 

%  "o 

.2   So 

*o  m 

alS 

2  ^  ^  1 

c  c 

0 

s  S 

c 

3 

c 

§    . 
•3  & 

c 

3  i 

1  p 

jp 

I*; 

^  2  "c 

>  H 

|2S| 

1^  § 

11 

£  -s 

c  1 

c   ^ 

s£ 

6  .s 

ii 

>  f_; 

6    2 
S 

in 

11 

^H 

—  Ijj 

Is 

m  3  o  'E 

W  o 

1 

£§ 

s 

i* 

lri 

OH 

SPQ 

&H 

*« 

1" 

I* 

ft  7  c 

&  £  a 

1 

3  *"  ^ 

1H 

S 

4) 

90 

0.00 

100.00 

8325 

8325 

8325 

1.68 

119 

2465° 

85 

28.33 

66^67 

5552 

3!33 

240 

1.67 

116 

5900 

339 

5561 

2.52 

119 

2236 

80 

42.50 

50.00 

4160 

5.00 

361 

2.50 

174 

4697 

509 

4188 

3.34 

120 

2023 

75 

51.00 

40.00 

3330 

6.00 

433 

3.00 

209 

3972 

611 

3361 

4.17 

120 

1862 

70 

56.67 

33.33 

2775 

6.67 

482 

3.33 

232 

3489 

679 

2810 

4.98 

120 

1732 

65 

60.71 

28.57 

2378 

7.15 

516 

3.57 

248 

3142 

727 

2415 

5.80 

121 

1612 

60 

63.75 

25.00 

2081 

7.50 

541 

3.75 

261 

2883 

764 

2119 

6.61 

121 

1513 

55 

66.12 

22.22 

1850 

7.78 

562 

3.88 

270 

2682 

792 

1890 

7.40 

121 

1427 

50 

68.00 

20.00 

1665 

8.00 

578 

4.00 

278 

2521 

815 

1706 

8.21 

122 

1350 

45 

69.55 

18.18 

1513 

8.18 

591 

4.09 

284 

2388 

833 

1555 

9.00 

122 

1284 

40 

70.83 

16.67 

1388 

8.33 

601 

4.17 

290 

2279 

849 

1430 

9.79 

123 

1222 

25 

73.67 

13.33 

1110 

8.67 

626 

4.33 

301 

2037 

883 

1154 

12.13 

124 

1077 

15 

75.00 

11.77 

980 

8.82 

637 

4.41 

307 

1924 

899 

1025 

13.66 

124 

1002 

0 

76.50 

10.00 

832 

9.00 

650 

4.50 

313 

1795 

916 

879 

15.93 

126 

906 

variation  with  the  degree  of  extraction.     A  typical  furnace  for  burning 
bagasse  is  shown  in  Fig.  4a. 

Bagasse  as  Fuel:  Prac.  Engr.  U.  S.,  Jan.,  1910;  Engr.  U.  S.,  April  1,  1903;  Jour. 
Assn.  Engng.  Soc.,  July,  1901;  Engng.,  Feb.  18,  1910. 

Tanbark  is  usually  quite  moist;  the  amount  of  moisture  varies  with 
the  leaching  process  used  and  averages  around  65  per  cent.  In  this 
condition  it  has  a  heat  value  of  about  4300  B.T.U.  per  pound.  If 
perfectly  dry  its  heating  power  is  approximately  6100  B.T.U.  per  pound. 
As  in  the  case  of  all  moist  fuels,  tanbark  must  be  surrounded  by  heated 
surfaces  of  sufficient  extent  to  insure  drying  out  the  fresh  fuel  as  thrown 
on  the  fire.  A  successful  furnace  for  burning  tanbark  is  shown  in 
Fig.  4b. 

Tanbark  as  a  Boiler  Fuel:  Jour.  A.S.M.E.,  Feb.,  1910,  p.  181;  Jour.  A.S.M.E., 
Oct.,  1909,  p.  951 ;  Prac.  Engr.  U.  S.,  Jan.,  1910. 


FUELS  AND  COMBUSTION 


21 


22 


STEAM  POWER  PLANT  ENGINEERING 


FUELS  AND  COMBUSTION 


23 


17.  Composition  of  Coal.  —  The  uncombined  carbon  in  coal  is  known 
as  fixed  carbon,  while  the  hydrocarbons  and  other  gaseous  compounds 
which  distill  off  on  application  of  heat  constitute  the  volatile  matter. 
Refractory  earths  and  moisture  are  found  in  varying  quantities  in 
different  classes  of  coal  and  as  they  are  incombustible  tend  to  reduce 
the  heat  value  of  the  fuel.  That  part  of  the  fuel  which  is  dry  and  free 
from  ash  is  called  the  combustible,  though  the  nitrogen  and  oxygen  in  the 
volatile  matter  are  not  actually  combustible.  The  term  "  pure  coal  " 
has  been  suggested  in  this  connection  and  is  meeting  with  much  favor. 
(Jour.  W.S.E.  11-757.)*  The  various  elementary  constituents  of  a 
fuel  must  be  determined  by  a  careful  chemical  analysis,  but  in  most 
cases  it  is  only  necessary  to  know  the  heating  value,  the  per  cent 
of  moisture  and  ash  and  perhaps  the  per  cent  of  sulphur.  Table  1 
shows  the  composition  of  a  number  of  American  coals  and  gives  a  good 
idea  of  their  chemical  characteristics. 


TABLE   1. 

COMPOSITION    OF   TYPICAL   AMERICAN    COALS. 
(U.S.  Geological  Survey.) 


Anthracite. 

Semi-Bituminous. 

| 

i#. 
|| 

Is 

2*** 

s*a 

II 

2  o 

-  c 

J3 
rt 

1  s 

I  i 

*. 

GQ     kT  ^ 

o 

ci  M 

g| 

•  3  J5 

a  w  53 

|WM 

Proximate  analysis 
Water 

1.97 
4.35 
86.49 
7.19 
0.64 

85.66 
2.78 
0.77 
2.87 

1.50 

7.84 
81.07 
9.59 
0.50 

83.20 
3.29 
0.95 
2.45 

2.08 
7.27 
74.32 
16.33 
0.77 

75.21 
2.81 
0.80 
4.08 

12472 
12395 

0.65 
18.80 
75.92 
4.63 
0.57 

85.91 
4.58 
1.07 
3.24 

15190 
15104 

0.59 
18.52 
74.31 
6.58 
0.81 

81.05 
4.91 
2.15 
4.57 

1.28 
12.82 
73.69 
12.21 
2.01 

77.29 
3.74 
1.39 
3.36 

13406 
13831 

Volatile  matter  .  .  . 
Fixed  carbon  
Ash         

Sulphur  
Ultimate  analysis 
Carbon         

Hvciroffen 

Nitrogen  

Oxygen    

Calorific  value 
Calorimeter 

Dulong's  Formula 

13963 

13954 

14484 

H.  J.  Williams. 


*  See  also  "  Unit  Coal  and  the  Composition  of  Coal  Ash,"  Univ.  of  111.  Bulletin 
No.  37,  Aug.  9,  1909. 


24 


STEAM  POWER  PLANT  ENGINEERING 
TABLE    I.  — Continued. 


Bituminous. 

Lignite. 

<f] 
11}  j 

bc^S  tf 
PQ 

a    r  d 

ill 

B  P^  ^ 
1 

1-9 

*l 

£  3  a 

tuO  <D  C 

fas 

OQ 

*|L- 

B  1  S 

°3£ 

% 

al 

ji 

1      § 
*8« 

^ll 
•3SS 

o 

Proximate  analysis 
Water  

8.61 
32.40 
51.33 
7.66 
1.65 

68.14 
5.38 
1.34 
15.83 

12236 
12082 

3.15 
30.27 
56.17 
10.41 
1.26 

74.33 
4.96 
1.43 
7.61 

13406 
13371 

1.92 
36.56 
57.08 
4.44 
1.24 

78.31 
5.36 
1.85 
8.80 

14319 
14081 

5.31 
34.29 
36.24 
24.16 
4.30 

54.06 
4.57 
0.78 
12.13 

9848 
9929 

18.51 
35.33 
30.67 
15.49 
3.05 

47.34 
5.93 
0.66 
27.53 

8525 
8550 

10.86 
35.14 
46.90 
7.10 
0.64 

64.34 
5.73 
1.05 
21.14 

11435 
11299 

Volatile  matter.  .  . 
Fixed  carbon 

Ash     . 

Sulphur  . 

Ultimate  analysis 
Carbon  

Hydrogen  
Nitrogen          .  . 

Oxygen  
Calorific  value 
Calorimeter  

Dulong's  Formula 

Report  of  Committee  on  Coal  Analysis:  Jour.  Amer.  Chem.  Soc.,  21-1116,  1899. 
Coal  Testing :  Mines  and  Min.,  Nov.,  1905.  Coal  Testing :  Col.  Guard,  April  12,  1906. 
Report  of  Government  Coal-Testing  Plant:  Trans.  A.S.M.E.,  Dec.,  1905;  Eng.  and 
Min.  Jour.,  Sept.  15,  1904;  Mines  and  Min.,  June,  1906.  Determination  of  Volatile 
Combustible  Matter:  Jour.  Am.  Chem.  Soc.,  21-1137,  1899;  ibid.,  28-1002,  1906. 
Proximate  Constituents  of  Coal:  Jour.  Gas  Light,  Jan.  5,  1897.  Ultimate  Analy- 
sis: Col.  Guard,  Oct.  15,  1897.  Analysis  of  Coal:  Trans.  A.S.M.E.,  21-65. 
Sampling  Coal:  Chem.  Engr.,  Nov.,  1905;  Am.  Inst.  Mech.  Engrs.,  Sept.,  1905. 
Illinois  Coal  Tests :  Eng.  News,  Feb.  7,  1907. 

Sulphur  in  Coal:  Eng.  and  Min.  Jour.,  June  27,  1903;  Mines  and  Min.,  Feb.,  1906; 
Jour.  Am.  Chem.  Soc.,  24-852,  25-184,  26-1139. 

Ash:  Relation  between  Composition  and  Fusibility  in  Coal  Ash:  Col.  Guard,  Oct., 
1897,  March  18,  1898.  Composition  of  Ash  :  Stromeyer,  Steam  Boilers,  p.  11. 

18.  Combustion.  —  By  combustion  is  meant  the  chemical  union  of 
the  combustible  material  of  a  fuel  and  the  oxygen  of  the  air.  Theoretic- 
ally the  process  is  a  simple  one,  as  it  is  only  necessary  to  bring  each 
particle  of  fuel  previously  heated  to  the  kindling  temperature  in  con- 
tact with  the  correct  amount  of  oxygen  and  the  combustion  will  be 
complete,  the  fuel  oxidizing  to  the  highest  possible  degree.  In  practice, 
however,  the  size  and  character  of  fuel,  type  of  furnace,  draft,  impuri- 
ties in  the  fuel,  and  the  mechanical  difficulties  affect  combustion  to 
such  an  extent  as  to  render  oxidation  more  or  less  incomplete. 

When  heat  is  applied  to  coal  combustion  takes  place  in  three 
separate  and  distinct  stages: 

1.  Absorption  of  heat.  A  fresh  charge  of  fuel  when  thrown  on  a 
fire  must  first  be  brought  to  the  kindling  point  in  order  that  chemical 


FUELS  AND  COMBUSTION  25 

action  may  take  place.     The  temperatures  necessary  to  cause  this  union 
of  oxygen  and  fuel  are  approximately  as  follows: 

Degrees  F.  Degrees  F. 

Lignite  dust 300     Cokes 800 

Sulphur 470     Anthracite  lump 750 

Dried  peat 435     Carbon  monoxide 1211 

Anthracite  dust 570     Hydrogen   1100 

Lump  coal 600 

(Stromeyer,  Marine  Boiler  Management  and  Construction,  p.  93.) 

2.  Vaporization    of    the  hydrocarbon  portion  of    the  fuel  and    its 
combustion,  the  hydrocarbons  consisting  principally  of    olefiant    gas, 
C2H2,  marsh  gas,  CH4,  tar,  pitch,  naphtha  and  the  like.     As  these  gases 
are  driven  off  they  become  mixed  with  the  entering  air,  and  the  carbon 
and  hydrogen  unite  with  the  oxygen,  forming  carbon  dioxide,  CO2,  and 
water  vapor,   H20,  respectively,   and  give  up  heat  in  doing  so.     If 
volatile  sulphur  is  present  it  unites  with  the  oxygen,  forming  sulphur 
dioxide,  SO2,  and  also  gives  up  heat,  but  ite  presence  is  objectionable,  as 
the  SO2,  particularly,  in  the  presence  of  moisture,  attacks  the  metal 
of  the  furnace  and  boiler  and  causes  rapid  corrosion.     If  insufficient 
oxygen  is  present  for  complete  oxidation,  the  carbon  may  burn  to  carbon 
monoxide,  CO,  and  only  a  small  portion  of  the  available  heat  be  liberated. 

3.  Combustion   of   the  solid  or  carbonaceous  portion   of   the   fuel. 
After  the  hydrocarbons  have  been  driven  off  and  oxidized  the  remain- 
ing solid  matter  is  composed  chiefly  of  carbon  and  ash.     The  carbon 
unites  with  the  oxygen,  forming  carbon  dioxide,  carbon  monoxide,  or 
both,  depending  upon  the  completeness  of  combustion.     The  ash,  of 
course,  remains  unconsumed. 

In  commercial  practice  the  requirements  for  perfect  combustion  are  a 
surplus  of  air,  a  thorough  mixture  of  the  fuel  particles  with  the  air,  and 
a  high  temperature.  The  surplus  of  air  above  theoretical  require- 
ments should  be  kept  to  a  minimum,  but  even  in  the  most  scientifically 
designed  furnace  some  excess  is  essential  on  account  of  the  difficulty 
of  properly  mixing  the  gases,  since  the  currents  of  combustible  gases  and 
air  are  apt  to  be  more  or  less  stratified.  The  products  of  combustion 
must  be  maintained  at  the  kindling  temperature  until  oxidation  is 
complete,  otherwise  the  carbon  will  be  wasted  as  carbon  monoxide  or 
as  smoke.  The  final  products  of  combustion  as  exhausted  by  the 
chimney  should  consist  only  of  carbon  dioxide,  water  vapor,  oxygen, 
nitrogen,  and  the  oxides  of  impurities  in  the  fuel. 

When  the  combustible  elements  unite  with  oxygen  they  do  so  in 
definite  proportions,  called  the  combining  weights,  which  are  always 
the  same,  and  the  union  produces  a  fixed  quantity  of  heat.  Thus  in  the 


STEAM  POWER  PLANT  ENGINEERING 


combustion  of  carbon  to  CO,  12  pounds  of  carbon  unite  with  16  pounds 
of  oxygen,  forming  28  pounds  of  CO;  hence  one  pound  of  carbon  will 


form 


12  +  16 
12 


2J  pounds  of  CO. 


The  heat  of  combustion  will  be  4,451  B.T.U.  per  pound  of  carbon  thus 

consumed. 

TABLE  2. 

DATA  RELATIVE  TO  ELEMENTS   MOST  COMMONLY   MET  WITH  IN  CONNECTION 
WITH    COMBUSTION    OF    FUEL. 


Substance. 

Chemical 
Formula. 

Combining 
Weight. 

Chemical  Reaction. 

Oxygen. 

Air. 

Weight  per  Pound 
Substance  in  First 
Column. 

Acetylene  

C2H2 

26 

C2H2+5O=2CO2  +  H2O 

3.08 

13.27 

Air  

Ash  

C 
C 
CO2 

CO 
H 

CNH< 
C6H. 

S 

H2O 

12 
12 
44 

28 
1 
16 
14 
28 
16 
32 
18 

C+0 
C+2 

ICO 

1.33 
2.66 

5.78 
11.58 

Carbon.    . 

0=C02 

Carbon  dioxide 
Carbon  monox- 
ide   

CO  +  0  =  C02 

2H  +  O  =  H20 
CH4+4O=C02+2H20 

0.57 

8 
4 

2.47 
34.8 
17.4 

Hydrogen  
Marsh  gas 

Nitrogen  

Olefiant  gas  .  .  . 
Oxygen  . 

C2H4+  6  O  =  2  CO2  +  2  H2O 

3.43 

14.9 

Sulphur  

S+2O=SO2 

1 

4.32 

Water  vapor  .  .  . 

Substance. 

Specific 
Heat  at 
Constant 
Pressure.  * 

Weight  per 
Cubic  Foot. 

Cubic  Feet 
per  Pound. 

Heat  of  Combustion 
B.T.U.t 

62°  F.        30"  Hg.J 

Per  Pound. 

Per  Cubic  Foot. 
62°  F.  30"  Hg. 

Acetylene  
Air  

6.2375 
0.2 

§ 

0.0685 
0.0759 

0.1159 

0.0737 
0.00527 
0.0421 
0.07376 
0.0737 
0.0843 
0.1686 

14.6 
13.16 

'  8*  .62  7 

13.55 
189.8 
23.72 
13.55 
13.55 
11.86 
5.93 

21,465 

4,451 
14,544 

4,325 
62,032 
23,513 

21,34*4 

1470 

320 
326 
991 

1*570 

Ash  

Carbon  
Carbon 

Carbon  dioxide 
Carbon  monox- 
ide   

0.2169 

0.2426 
3.409 
0.593 
0.2438 
0.404 
0.2175 
0.1544 
0.48 

Hydrogen  
Marsh  gas  
Nitrogen 

Olefiant  gas  .  .  . 
Oxvsren 

Sulphur  . 

Water  vapor..  . 

Regnault. 


t  Favre  &  Silberman. 


J  Smithsonian  Tables. 


§SOa. 


FUELS  AND  COMBUSTION  27 

Similarly  in  burning  to  CO2  one  pound  of  carbon  will  form  3f  pounds 
of  CO,  and  liberate  14,544  B.T.U. 

Table  2  gives  the  physical  and  chemical  properties  of  the  substances 
most  commonly  met  with  in  connection  with  combustion. 

Combustion:  Engr.  U.  S.,  Feb.  16,  1903,  April  1,  1903,  Jan.  1,  1907;  Power,  Feb. 
1905,  Oct.  26,  1906;  Eng.  Mag.,  July,  1907;  Prac.  Engr.  U.  S.,  Jan.,  1910. 

Simple  Method  of  Determining  Condition  of  Combustion  :  A.  Bement,  Jour.  West. 
Soc.  Eng.,  June,  1901.  Effect  of  Altitude  on  Combustion  :  Power,  Sept.,  1906.  Notes 
on  Fuel  Combustion  in  Power  Plants  :  Elec.  Engr.,  Lond.,  Aug.  10,  1906. 

19.  Temperature  due  to  Combustion.  —  If  the  heat  liberated  by  the 
chemical  union  of  any  two  elements  is  confined  in  such  a  way  that  no 
radiation  can  take  place,  the  resulting  increase  in  temperature  of  the 
products  of  combustion  may  be  expressed 


—,  a) 

ws 


in  which 

t  =  the  increase  in  temperature,  degrees  F. 
h  =  the  heat  liberated,  B.T.U.  per  pound  of  combustible. 

s  =  the  specific  heat  of  the  resulting  gaseous  products. 
w  =  weight  of  the  gaseous  products,  pounds. 

Thus  1  pound  of  carbon  in  burning  to  CO2  combines  with  2f  pounds 
of  oxygen  and  forms  3§  pounds  of  CO2.  Taking  the  mean  specific 
heat  of  C02  as  0.217,  the  resulting  increase  in  temperature  will  be 


Such  a  temperature,  of  course,  cannot  be  reached  in  practice,  even 
with  pure  carbon  and  oxygen,  on  account  of  the  radiation  losses. 

In  the  ordinary  furnace  the  oxygen  is  obtained  from  the  atmosphere, 
which,  neglecting  a  few  impurities  and  minor  elements,  contains  the 
following,  mechanically  mixed: 

Oxygen  ......................................  23  parts  by  weight. 

Nitrogen  ...................................  77  parts  by  weight. 

or 

Oxygen  .....................................  21  parts  by  volume. 

Nitrogen  ..............  ......................  79  parts  by  volume. 


Hence  in  the  combustion  of  one  pound  of  pure  carbon  the  products 
of  combustion  contain  not  only  3f  pounds  of  C02,  but  —  X  2§  =  8.92 

Zo 

pounds   of  nitrogen,    giving   a    total    of   3|  +  8.92  =  12.58    pounds. 


28  STEAM    POWER    PLANT    ENGINEERING 

The  nitrogen  performs  no  useful  office  in  combustion  and  passes  through 
the  furnace  without  change.  It  dilutes  the  products  of  combustion  and 
reduces  their  temperature.  Thus  in  the  specific  case  taken  above  the 
theoretical  increase  in  temperature  will  be 

14  f\AA 
,     *  -  12.58  X  0.236  =  49°°  ^^  R  ;  '       .V 

The  mean  specific  heat,  0.236,  of  the  products  of  combustion  is  determined  as 
follows: 

Heat  necessary  to  raise  8.92  Ibs.  of  nitrogen  1  degree  is  8.92  X  0.2438  (specific 
heat  of  nitrogen)  =  2.1747  B.T.U. 

Heat  necessary  to  raise  3.66  Ibs.  of  CO2  1  degree  is  3.66  X  0.217  (specific  heat  of 
CO2)=  0.7942  B.T.U. 

Total  heat  =  2.1747  +  0.7942  =  2.9689  B.T.U. 

2  Qfi&Q 

Mean  specific  heat  =  ^£r         ^  =  0.236. 
o.oo 


Evidently  for  maximum  temperature  this  dilution  should  be  kept 
as  low  as  possible.  Unfortunately,  in  practice,  a  perfect  union  of  fuel 
and  air  in  theoretical  proportions  is  almost  impossible,  and  to  insure 
complete  combustion  an  excess  of  air  is  necessary.  This  still  further 
reduces  the  temperature  of  the  products  of  combustion.  For  example, 
the  complete  oxidation  of  one  pound  of  carbon  requires  11.5  pounds  of 
air.  If  20  pounds  of  air  are  supplied  per  pound  of  carbon,  an  average 
figure  in  good  practice,  the  theoretical  increase  in  temperature  will  be 

14  ^44 

-  2934 


•  200236 

In  the  preceding  computation  the  specific  heat  of  the  constituent 
gases  is  assumed  to  be  constant  for  all  temperatures.  Recent  experi- 
ment seems  to  show  that  the  specific  heat  increases  at  high  tempera- 
tures, though  actual  values  are  not  yet  available. 

20.  Air  required  for  Combustion.  —  The  quantity  of  air  required 
for  complete  combustion  may  be  approximately  determined  from  the 
ultimate  analysis.  Thus 

A  =  34.56^+H-^,  (2) 

in  which 

A  =  weight  of  air  required  per  pound  of  fuel. 

C,  H,  and  O  =  proportional  part  by  weight  of  carbon,  hydrogen,  and 
oxygen  in  the  fuel. 

When  the  flue-gas  analysis  is  known,  the  air  actually  supplied  may  be 
approximately  determined  as  in  the  following  example: 

Example:  Required  the  weight  of  air  supplied  per  pound  of  coal  with 
fuel  and  flue  gas  analysis  as  follows: 


FUELS  AND  COMBUSTION 


29 


Coal  analysis  (Illinois  bituminous,  mine  run): 

Carbon  68.14  Nitrogen  1.34 

Oxygen  15.83  Ash  7.66 

Hydrogen          5.38  Sulphur  1.65 

Flue  gas  analysis,  cubic  feet  per  100  cubic  feet  of  dry  gas  (tempera- 
ture, 62°  F.,  barometer  30  inches): 
CO2      14 
CO         0.5 
0         6.0 
S02       (not  determined,  since  its  weight  is 

practically  negligible). 
The  weights  may  be  determined  from  the  densities  given  in  Table  2. 


Volume. 

Density. 

Weight. 

CO2 

o 

CO 

14 
6 
0.5 

0.1159 
0.0843 
0.0737 

1.6226 
0.5058 
0.0368 

These  weights  can  be  subdivided  into  those  of  their  constituents; 
thus  the  CO2  contains  f\  of  carbon  and  T8T  of  oxygen,  and  the  CO,  f  of 
carbon  and  f  of  oxygen. 

T8T  X  1.6226  =  1.1800  T8T  X  1.6226  =  0.4426 

f  X  0.0368  =  0.0210  f  X  0.0368  =  0.0158 

0.5058  Pounds  of  carbon  0.4584 


I 

' 


0.4584 


3.72 


=  3.72. 


Pounds  of  oxygen  1.7068 
Weight  of  oxygen  per  pound  of  carbon  = 
Since  23%  of  air  by  weight  is  oxygen, 

Weight  of  air  per  pound  of  carbon  =  TT^  —  16.2. 

Weight  of  air  supplied  per  pound  of  coal  is 
for  the  carbon, 

for  the  hydrogen,  34.56^0.0538  -  °'1^83^=   1<17  lbs> 
total  N  '     12.21  Ibs. 

The  theoretical  weight  of  air  for  combustion  would  have  been 


0.6814  X  16.2  =  11.04  Ibs. 


11.58  X  0.6814+34 


Wo. 


0538  - 


=  9.o6 


The  percentage  of  air  excess  =  100 


-  9.06 


9.06 


The  weight  of  air  actually  supplied  may  be  closely  approximated 
by  the  following  equation : 

/  AT  \ 

:  C,  (2a) 


30 


STEAM  POWER  PLANT  ENGINEERING 


in  which     A  =  the  weight  of  air  supplied  per  pound  of  fuel. 
N,  CO,  CO2  =  percentages  by  volume  of  nitrogen, 

carbon  monoxide,  and  carbon  dioxide  in  the  flue  gas. 
C  =  the  proportional  part  by  weight  of  carbon  in  the  fuel. 
The  per  cent  of  excess  air  supplied  per  pound  of  fuel  may  be  con- 
veniently determined  from  the  relationship 

Air  actually  required       _  N 

Air  theoretically  supplied  ~  N  -  3.782  O* 

N  and  O  are  respectively,  by  volume,  the  proportional  parts  of  the 
nitrogen  and  oxygen  in  the  flue  gas.  The  free  oxygen  is  due  to  the  air 
supplied  and  not  used.  This  oxygen  was  accompanied  by  3.782  times 
its  volume  of  nitrogen.  (N  —  3.782)  represents  the  nitrogen  content 
in  the  air  actually  required  for  combustion.  Hence,  N-v-(N  —  3.782) 
is  the  ratio  of  the  air  supplied  to  that  required. 

Table  2a  gives  the  values  of  the  ratio  corresponding  to  various  per- 
centages of  CO2  +  CO  and  CO2  +  CO  +  O. 

TABLE   2a. 

RATIO  OF  TOTAL  AIR  SUPPLIED  TO  THAT  THEORETICALLY  REQUIRED 
FOR   VARIOUS  ANALYSES  OF  FLUE-GASES. 

N 


Ratio  = 


N-  3.782O 


CO2+CO. 

N=79. 
C02+C0 
+  0=21. 

N  =  79.5. 
CO2+CO 
+O=20.5. 

N=80. 
CO2  +  CO 
+  O=20. 

N=80.5. 
CO2  +  CO 
+O=19.5. 

N=81. 
CO2  +  CO 
+  O=19. 

N=81.5. 
CO2+CO 
+O=18.5. 

N=82. 
CO2  +  CO 
+  0=18. 

21 

.02 

20 

.05 

1.00 

1.00 

19 

11 

1.08 

1.05 

.02 

.00 

18 

'.17 

.14 

.10 

.08 

.05 

1.02 

.00 

17 

.24 

.20 

.17 

.13 

.10 

1.07 

.05 

16 

.32 

.27 

.23 

.20 

.16 

1.13 

.10 

15 

.40 

.35 

.31 

.27 

.23 

1.19 

.16 

14 

.51 

.45 

.39 

.35 

.30 

1.26 

.23 

13 

.62 

1.55 

.50 

.44 

.39 

1.34 

.30 

12 

.76 

1.68 

.61 

.54 

.49 

1.43 

.38 

11 

.92 

1.82 

.74 

.66 

.60 

1.53 

.48 

10 

2.11 

2.00 

1.90 

1.81 

.72 

1.65 

.59 

9 

2.35 

2.21 

2.08 

1.97 

.88 

1.79 

1.71 

8 

2.65 

2.47 

2.31 

2.18 

2.06 

1.95 

1.86 

7 

3.03 

2.80 

2.59 

2.44 

2.27 

2.14 

2.03 

6 

3.55 

3.22 

2.96 

2.74 

2.54 

2.38 

2.24 

5 

4.27 

3.81 

3.44 

3.14 

2.89 

2.68 

2.50 

4 

5.37 

4.65 

4.11 

3.68 

3.34 

3.05 

2.83 

3 

7.23 

5.97 

5.10 

4.45 

3.96 

3.56 

3.25 

2 

11.06 

8.34 

6.71 

5.63 

4.85 

4.27 

3.82 

1 

23.51 

13.83 

9.83 

7.64 

6.27 

6.12 

4.64 

The  weight  of  air  per  pound  of  combustible  as  indicated  by  the  per 
cent  of  CO2  in  the  flue  gas  is  given  in  Table  3.  These  figures  are  only 
approximate  but  are  sufficiently  accurate  for  many  practical  purposes. 


FUELS  AND  COMBUSTION 


31 


TABLE  3. 

WEIGHT  OF  AIR  PER  POUND  OF  COMBUSTIBLE  AS  INDICATED  BY  THE  PER- 
CENTAGE   OF   C02   IN   THE    FLUE   GAS. 


Per  Cent  of  CO2. 

Pounds  of 
Air. 

Per  Cent  of 
C02. 

Pounds  of 
Air. 

Per  Cent  of 
C02. 

Pounds  of 
Air. 

21 

12 

14 

18 

.   7 

36 

20 

12.6 

13 

19.4 

6 

42 

19 

13.3 

12 

21 

5 

50.5 

18 

14 

11 

22.9 

4 

63 

17 

14.8 

10 

25.2 

3 

84 

16 

15.7 

9 

28 

2 

126 

15 

16.8 

8 

31.5 

1 

210 

21.  Calorific  Value  of  Coal.  —  The  heat  liberated  by  the  combustion 
of  unit  weight  of  fuel  is  called  the  calorific  value  of  the  fuel.  The  most 
rational  way  of  determining  the  heat  of  combustion  is  to  burn  a 
weighed  sample  of  coal  in  an  atmosphere  of  oxygen  in  a  suitable 
calorimeter.  An  alternative  method  is  to  calculate  the  heat  of  com- 
bustion from  the  chemical  analysis.  An  analysis  which  determines 
the  per  cent  of  fixed  carbon,  volatile  matter,  moisture,  and  ash  is  called 
the  proximate  analysis,  while  one  which  reduces  the  fuel  to  its  ele- 
mentary constituents  of  carbon,  hydrogen,  nitrogen,  sulphur,  moisture, 
and  ash  is  called  the  ultimate  analysis.  The  proximate  analysis  is 
comparatively  easy  to  make  and  gives  the  general  characteristics  of  a 
fuel.  It  is  made  by  subjecting  the  sample  to  a  moderate  temperature 
to  expel  the  moisture,  then  to  a  higher  temperature  until  the  volatile 
matter  is  driven  off,  and  finally  to  a  very  high  temperature  which  drives 
off  all  carbon  as  carbon  dioxide  and  leaves  the  ash  as  a  residue.  By 
weighing  the  residue  at  the  end  of  each  operation  the  various  percent- 
ages may  be  computed.  For  method  of  making  proximate  analysis,  see 
"  Report  of  Committee  on  Coal  Testing,"  Journal  of  the  American 
Chemical  Society,  Vol.  21,  p.  1116. 

The  heating  value  of  certain  classes  of  coals  may  be  estimated  from 
the  proximate  analysis.  Thus  for  Illinois  coals  with  ash  content  under 
10  per  cent,  R.  W.  Hunt  &  Co.  deduced  the  formula 


h  =  14,544  C  +  16,515  V  -  10,000  A, 


(3) 


in  which 


h  =  B.T.U.  per  pound  of  coal. 

C,  V,  and  A  =  the  proportional  content  of  fixed  carbon,  volatile 
matter,  and  ash. 


32  STEAM  POWER  PLANT  ENGINEERING 

When  ash  lies  between  10  and  15  per  cent,  the  formula  will  be  more 
accurate  if  written 

h  =  14,544  C  +  16,515  V  +  354  A  -  1635  (4) 

Kent,  Goutal,  and  other  authorities  have  deduced  formulas  based  on 
the  proximate  analysis  which  agree  closely  with  the  calorimetric  deter- 
minations for  many  coals,  but  a  calorimetric  determination  is  necessary 
whenever  exact  results  are  required. 

The  formula  most  commonly  used  in  calculating  the  heating  value  of 
a  fuel  is  based  on  the  ultimate  analysis  and  is  known  as  Dulong's 
formula.  Thus 

h  =  14,600  C  +  62,000  f H  -  ^ +  4,000  S,  (5) 

in  which 

h  —  heating  value  in  B.T.U.  per  pound  of  coal. 
C,  H,  O,  and  S  refer  to  the  proportions  of  carbon,  hydrogen,  oxygen, 
and  sulphur,  respectively. 

The  calorimetric  determination  is  more  readily  made  and  is  usually 
more  reliable  than  the  calculation  from  the  ultimate  analysis.  Table  1 
gives  the  proximate  and  ultimate  analyses  and  the  calorimetric  and  cal- 
culated heat  values  for  a  number  of  American  coals. 

Calorific  Value  of  Fuels:  Poole,  Engr.  U.S.,  April  1,  1903;  P.  Mahler,  Jour. 
Frank.  Inst.,  Jan.,  1905;  A.  Adams,  Jour.  Soc.  Chem.  Ind.,  Oct.  31  and  Nov.  30, 
1901;  Power,  Oct.  1,  1906;  Mines  and  Min.,  May,  1902;  Trans.  Am.  Inst.  Min.  Engrs., 
27-259;  Kent,  Steam  Boiler  Economy,  Chap.  V;  Report  of  Government  Coal  Testing 
Plant  at  St.  Louis. 

Calorimetry :  Berthier  Method  of  Coal  Calorimetry :  Trans.  A.S.M.E.,  21-304. 
Comparison  of  Calorimeters :  Jour.  Soc.  Chem.  Ind.,  22-1230,  1903 ;  23-704.  1m- 
proved  Form  of  Thompson  Calorimeter :  Jour.  Soc.  Chem.  Ind.,  25-409,  1906.  Parr's 
Calorimeter :  Chem.  Engr.,  Feb.,  1907.  Coal  Calorimeters :  Trans.  A.S.M.E.,  14- 
816,  16-1040.  Mahler  Bomb:  Calorific  Power  of  Fuels,  Poole;  Trans.  Am.  Inst. 
Min.  Engrs.,  27-259,  Engr.  U.S.,  Jan.  1,  1907,  p.  68. 

22.  Heat  Losses  in  Burning  Coal.  —  The  function  of  the  boiler  is  the 
absorption  of  the  heat  liberated  by  the  combustion  of  the  fuel.  In 
practice  from  50  per  cent  to  85  per  cent  of  this  heat  is  utilized  in  mak- 
ing steam,  depending  upon  the  conditions  of  operation,  and  the'  remainder 
is  wasted.  Complete  utilization  of  the  heat  of  combustion  is  impossible. 
A  boiler  and  furnace  which  utilizes  80  per  cent  of  the  heat  in  the  fuel 
is  exceptional  and  an  average  figure  for  very  good  practice  is  not 


FUELS  AND  COMBUSTION  '33 

far  from  75  per  cent.     The  various  losses  may  be  summed  up  as 
follows: 

1 .  Loss  in  the  dry  chimney  gases. 

2.  Loss  due  to  incomplete  combustion. 

3.  Loss  of  fuel  through  the  grate. 

4.  Superheating  the  hygroscopic  moisture  in  the  air. 

5.  Moisture  in  the  fuel. 

6.  Loss  due  to  the  presence  of  hydrogen  in  the  fuel. 

7.  Unburned  fuel   carried  beyond  the  combustion  chamber  in  the 
form  of  soot  or  smoke. 

8.  Radiation  and  minor  losses. 

Fuel  Losses:  G.  H.  Barrus,  Cassier's  Mag.,  Aug.,  1907;  Power,  Sept.  14, 1909,  p.  439. 

23.  Loss  in  the  Dry  Chimney  Gases.  —  The  most  serious  loss  is  the 
heat  carried  away  in  the  chimney  gases.  This  may  be  expressed 

h  =  (W  +  1)  (T  -  t)  S,  (6) 

in  which 

h  =  loss  in  B.T.U.  per  pound  of  combustible. 
W  =  weight  of  air  supplied  per  pound  of  combustible. 
T   =  chimney  temperature,  degrees  F. 

i   =  temperature  of  the  air  entering  furnace. 

S  =  specific  heat  of  the  gases.     (This  may  be  taken  as  0.24  for 
most  purposes.) 

This  loss  is  unavoidable  and  is  seldom  less  than  15  per  cent  of  the  total 
heat  supplied.  In  average  good  practice  it  is  not  far  from  20  per  cent. 
Even  in  the  ideal  case  when  pure  carbon  burns  to  CO2  and  the  air 
supply  is  theoretically  correct,  the  loss  is  appreciable. 

For  example,  suppose  the  temperature  of  the  air  is  50  degrees  F.,  the 
temperature  of  the  flue  gas  450  degrees  F.,  and  the  air  supply  11.5 
pounds  per  pound  of  combustible,  the  heat  carried  away  by  the  chim- 
ney will  be 

12.5  X  (450  -  50)  X  0.24  =  1200  B.T.U., 

which  is  1200  -*-  14,500  =  8.3  per  cent  of  the  total  heat  supplied. 

As  a  matter  of  fact,  considerable  excess  of  air  is  necessary  for  com- 
plete combustion,  depending  upon  the  size  and  nature  of  the  fuel, 
thickness  of  the  fire,  variation  in  resistance  through  the  fuel,  intensity 
of  draft,  and  character  of  grate.  Practice  shows  a  minimum  excess  of 
50  per  cent  in  well-proportioned  furnaces  working  under  ideal  conditions 
with  an  average  not  far  from  100  per  cent. 

Table  4  indicates  the  magnitude  of  these  losses  for  different  chimney 
temperatures  and  weights  of  air  per  pound  of  combustible. 


STEAM  POWER  PLANT  ENGINEERING 


TABLE  4. 

HEAT  CARRIED  AWAY  BY  THE  DRY  CHIMNEY  GASES  PER  POUND  OF 

COMBUSTIBLE. 


Pounds  of  Air  per  Pound  of  Combustible. 

Temperature  of  Chimney  Gases.  Deg.  Fahr. 

300° 

350° 

400° 

450° 

500° 

550° 

600° 

650° 

12 
# 

750 

5.2 

905 

6.2 

1060 

7.3 

1216 

8.7 

1370 

9.5 

1528 

10.5 

1683 

11.6 

1840 

12.7 

15 

865 

6 

1112 

7.6 

1305 

9.1 

1498 

10.3 

1679 

11.6 

1880 
13.0 

2072 

14.3 

2262 

15.6 

18 

1004 

7.2 

1321 

9.1 

1550 

10.7 

1778 
12.2 

2010 

13.9 

2235 

15.4 

2460 

17 

2692 

17.9 

21 

1266 

8.7 

1530 

10.5 

1785 

12.3 

2060 

14.2 

2320 

16 

2582 

17.8 

2846 

19.5 

3118 

21 

24 
27 
30 
33 

1440 

9.9 

1740 

12 

2040 

14 

2340 

16.1 

2640 

18.2 

2940 

20.3 

3240 

22.4 

3540 

24.4 

1611 

11.1 

1950 

13.5 

2281 

15.7 

2620 

18.1 

2958 

20.4 

3291 

22.7 

3628 

25 

3962 

27.4 

1785 

12.4 

2160 

14.9 

2530 

17.4 

2900 

20 

3270 

22.6 

3641 

25 

4016 

27.8 

4396 

30.4 

1957 

13.5 

2362 

16.3 

2779 

19.2 

3180 

22 

3589 

24.7 

4000 

27.6 

4405 

30.5 

4820 

33.2 

36 

2130 

14.7 

2579 

17.8 

3020 

20.8 

3461 

23.9 

3910 

27 

4350 

30 

4798 

33 

5290 

36.6 

39 

2300 

15.9 

2781 

19.2 

3261 

22.5 

3743 

25.8 

4220 

29.2 

4700 

32.4 

5180 

35.7 

5670 

39 

42 

2479 

17.1 

2999 

20.6 

3508 

24.7 

4023 

27.7 

4540 

31.3 

5052 

34.8 

5570 

39.4 

6100 

42 

*  Theoretical  requirement. 

Large  type  gives  the  loss  in  B.T.U.  per  pound  of  combustible. 
Small  type  gives  the  per  cent  loss,  assuming  a  calorific  value  of  14,500  B.T.U. 
per  pound  of  combustible. 

From  the  table  it  will  be  seen  that  this  loss  is  approximately  12  per 
cent  of  the  total  heat  supplied  when  the  excess  of  air  is  50  per  cent  and 
16  per  cent  when  the  excess  is  100  per  cent.  Flue-gas  temperature 
assumed  to  be  450  degrees  F.  Flue  temperatures  less  than  450  degrees  F. 
are  seldom  experienced  except  in  connection  with  economizers,  and  an 
average  figure  is  about  500  degrees  F. 


FUELS  AND  COMBUSTION 


35 


The  weight  of  air  supplied  per  pound  of  combustible  is  most  con- 
veniently determined  by  the  percentage  of  CO2  in  the  flue  gas.  Thus 
for  the  complete  oxidation  of  pure  carbon  the  resulting  flue  gases 

TABLE  5. 

LOSS    DUE    TO    INCOMPLETE    COMBUSTION    OF    CARBON    TO    CARBON 

MONOXIDE. 


o3 
'   B 

3 

Per  Cent  of  CO2  in  the  Flue  Gas  by  Volume. 

6 

8 

10 

12 

14 

16 

0.2 

328 

2.2 

248 

1.7 

199 

1.3 

168 

1.1 

144 

1 

126 

0.8 

0.4 

635 

4.3 

484 

3.3 

390 

2.6 

327 

2.2 

282 
1.9 

248 

1.7 

Per  Cent  of  CO  in  the  Flue  Gas  by  Vo 

0.6 

925 

6.3 

709 

4.8 

575 

3.9 

474 

3.2 

417 

2.8 

367 

2.5 

0.8 

1192 

8.1 

923 

6.3 

750 

5.1 

635 

4.3 

549 

3.7 

495 

3.4 

1.0 

1494 

10.2 

1128 

7.7 

923 

6.3 

780 

5.3 

676 

4.6 

596 

4.1 

1.2 

1690 

11.5 

1321 

9 

1085 

7.4 

923 

6.3 

801 

5.4 

708 

4.8 

1.4 

1920 

13.1 

1512 

10.3 

1248 

8.5 

1061 

7.2 

924 

6.3 

819 

5.6 

1.6 

2104 

14.3 

1693 

11.5 

1400 

9.5 

1193 

8.1 

1040 

7.1 

924 

6.3 

1.8 

2340 

16 

1865 

12.7 

1549 

10.5 

1321 

9.0 

1151 

7.8 

1025 

7 

2.0 

2537 

17.2 

2030 

13.8 

1690 

11.5 

1450 

9.9 

1270 

8.6 

1129 

7.7 

Large  type  gives  the  loss  in  B.T.U.  per  pound  of  carbon.     Small  type  gives  the 
per  cent  loss,  assuming  a  calorific  value  of  14,650  B.T.U.  per  pound  of  carbon. 

should  consist  of  carbon  dioxide  and  nitrogen  only,  and  in  the  ratio 
by  volume  of  21  to  79;  therefore  21  per  cent  of  CO2  in  the  flue  gas 
is  indicative  of  complete  combustion  and  theoretical  air  supply.  In 
other  words,  the  ratio  by  volume  of  CO2  to  N  after  complete  com- 
bustion is  practically  the  same  as  the  ratio  of  the  oxygen  to  the  nitrogen 
in  the  air  before  combustion.  Table  3  gives  the  approximate  weight 


36 


STEAM  POWER  PLANT  ENGINEERING 


of  air  used  per  pound  of  combustible  for  different  percentages  of  CO2 
in  the  flue  gas. 

In  practice,  15  per  cent  is  all  that  can  be  expected  under  the  best 
conditions,  with  an  average  between  10  per  cent  and  12  per  cent.  Any- 
thing less  than  10  per  cent  shows  an  excessive  amount  of  air  supplied. 
Traveling  grates,  unless  carefully  operated,  are  apt  to  show  as  low  as 
5  per  cent  of  CO2. 

24.  Loss  due  to  Incomplete  Combustion.  —  If  the  volatile  gases  are 
not  completely  oxidized,  as  when  the  air  supply  is  insufficient  or  the 
mixture  of  air  and  gases  is  not  thorough,  some  of  the  carbon  may 
escape  as  CO.  The  presence  of  even  a  small  amount  of  CO  in  the  flue 
gas  is  indicative  of  a  very  appreciable  loss,  as  will  be  seen  from  Table  5. 
Carbon  monoxide  is  a  colorless  gas,  and  its  presence  in  the  chimney 
gases  cannot  be  detected  by  the  fireman,  consequently  the  absence  of 
smoke  is  not  an  infallible  guide  for  perfect  combustion.  This  loss  may 
be  expressed 


=  C  X 


10,150  CO 
(C02  +  CO) ' 


(7) 


in  which  h2  =  the  loss  in  B.T.U.  per  pound  of  carbon,  CO2  and  CO 
are  percentages  by  volume  of  the  flue  gases  and  C  is  the  proportional 
part  of  carbon  in  the  combustible. 


15 

«  14 

a  13 

|  12 

t  11 

I  10 

g  9 

'—  o 

*  ? 


Relation  of  Gas  Composition  in  Bear 

Combustion  Chamber  To  Temperature 

at  Same  Place 


1900       2000 


2100         2200         2300         2400         2500         2600 
Combustion  Chamber  Temperature.Deg.Fah. 


2700 


2800 


0.5 
0.4 

0.8  £ 
0.20 

0.1 
I 


FIG.  5.     Relation  of  Gas  Composition  in  Combustion  Chamber  to  Temperature. 

This  loss,  however,  may  be  wholly  avoided  in  a  properly  designed 
and  carefully  operated  furnace.  In  fact  the  loss  from  this  cause  is 
often  exaggerated  and  seldom  exceeds  2  per  cent  of  the  total  heat  value 
of  the  fuel  except  during  the  few  moments  following  the  replenishing 


FUELS  AND  COMBUSTION  37 

of  a  burned-down  fire  with  fresh  fuel  or  when  the  supply  of  air  is  checked 
to  meet  a  sudden  reduction  in  load.  In  improperly  designed  furnaces 
in  which  the  volatile  gases  are  brought  into  contact  with  the  cooler 
boiler  surface  before  combustion  is  complete,  the  carbon  monoxide  may 
be  reduced  in  temperature  below  its  ignition  point  and  consequently 
will  fail  to  combine  with  the  oxygen.  In  such  a  case  the  loss  may 
prove  to  be  a  serious  one.  Fig.  5  shows  the  relation  between  the  com- 
position of  the  products  of  combustion  in  -the  rear  combustion  chamber 
of  a  250-horse-power  Heine  boiler,  hand  fired,  and  the  temperature  at  the 
same  place.  (For  an  extended  discussion  of  this  subject  see  Jour. 
West.  Soc.  Engrs.,  June,  1907,  p.  285.) 

25.  Loss  of   Fuel  through  Grate.  —  The  refuse  from  a  fuel  is  that 
portion  which  falls  into  the  pit  in  the  form  of  ashes,  unburned  or  partially 
burned  fuel,  and  cinders.     The  loss  from  this  cause  depends  upon  the 
size  of  the  fuel,  the  width  of  opening  in  the  grate  bars,  and  the  type  of 
grate.     Coal  which  necessitates  frequent  slicing  is  apt  to  give  greater 
loss  than  a  free-burning  coal.     Under  good  conditions  of  operation  it 
ought  not  to  exceed  2  per   cent  of  the  total   heat  value  of  the  fuel. 
In  traveling  grates  in  which  a  large  percentage  of  the  fine  fuel  falls 
through  the  front  end  of  the  grate  a  special  hopper  is  ordinarily  installed 
in  the  ash  pit  which  reclaims  most  of  it.     (See  Fig.  99.) 

Loss  of  Fuel  in  Ashes  :  Power,  March,  1905.  Experiments  on  Fuel  Value  of  Bitumi- 
nous Coal  Ashes:  Technology  Quarterly,  Dec.,  1905.  Coal  Ash:  Jour.  Soc.  Chem. 
Ind.,  Jan.  15,  1904. 

26.  Superheating  the  Moisture  in  the  Air.  —  The  loss  due  to  this 
cause  is  a  minor  one,  though  on  hot,  humid  days  it  may  be  appreciable. 
This  loss  may  be  expressed 


-t)*  (8) 

in  which      h3  =  B.T.U.  lost  per  pound  of  combustible. 

M  =  weight    of    moisture    introduced    with    the    air    per 

pound  of  combustible. 

t  =  temperature  of  air  entering  the  furnace,  degrees  F. 
T  =  temperature  of  chimney  gases,  degrees  F. 

27.  Moisture  in  the  Fuel.  —  Moisture  in  the  fuel  represents  an  appre- 
ciable loss  in  economy  if  present  in  large  quantities,  since  the  heat 
necessary  to  evaporate  it  into  superheated  steam  at  chimney  temperature 
is  lost.  Firemen  occasionally  wet  the  coal  to  assist  coking  or  to  reduce 

*  The  latest  accepted  value  for  the  mean  specific  heat  of  water  vapor  at  atmos- 
pheric pressure  is  0.46  in  place  of  0.48. 


38          STEAM  POWER  PLANT  ENGINEERING 

the  dust,  but  moisture  thus  added  necessarily  reduces  the  furnace 
efficiency.     The  loss  due  to  this  cause  is  expressed: 

h4  =  M  [(212  -  0  +  966*  +  0.48f  (T  -  212)],  (9) 

in  which  h4  =  B.T.U.  lost  per  pound  of  combustible. 

M  =  weight  of  moisture  per  pound  of  combustible. 

Other  notations  as  in  preceding  equation. 

For  example,  the  heat  loss  due  to  the  moisture  in  a  pound  of  fuel 
containing  10  per  cent  water,  temperature  of  fuel  80  degrees  F.,  chimney 
temperature  480  degrees  F.,  is 

ht  =  0.1  [(212  -  80)  +  966*  +  0.48f  (480  -  212)] 
=  122.6  B.T.U. 

A  rough  rule  is  to  allow  a  loss  of  1  per  cent  of  the  total  heat  value  of 
the  dry  fuel  for  each  10  per  cent  of  moisture  present. 

28.  Loss  due  to  the  Presence  of  Hydrogen  in  the  Fuel.  —  The  hydro- 
gen in  any  fuel  which  is  not  rendered  inert  by  oxygen  burns  to  water 
and  in  so  doing  liberates  62,032  B.T.U.  per  pound.     All  of  this  heat  is 
not  available  for  producing  steam  in  the  boiler,  since  the  water  formed 
by  combustion  is  discharged  with  the  flue  gases  as  superheated  steam 
at  chimney  temperature.     This  loss  is  equal  to 

hb  =  9H  [(212  -  0  +  966*  +  0.48f  (T  -  212)],  (10) 

in  which  hb  =  B.T.U.  lost  per  pound  of  combustible. 

H  =  weight  of  hydrogen  per  pound  of  combustible. 

All  other  notations  as  in  equations  (8)  and  (9). 

With  anthracite  coal  this  loss  is  approximately  2.5  per  cent  of  the 
total  heat  value  of  the  combustible  and  with  bituminous  coal  it  runs  as 
high  as  4.5  per  cent. 

29.  Loss  due  to  Smoke.  —  Visible  smoke  consists   of   carbon  in  a 
flocculent  state  mixed  with  the  products  of  combustion.     It  is  seldom 
evident  in  connection  with  anthracite  coal  and  is  generally  associated 
with  bituminous  fuel.     A  smoky  chimney  does  not  necessarily  indicate 
an  inefficient  furnace,  since  the  losses  due  to  visible  smoke  generation 
seldom  exceed  2  per  cent;  as  a  matter  of  fact,  a  smoky  chimney  may 
be  much  more  economical  than  one  which  is  smokeless.     That  is  to 
say,  a  furnace  operating  with  minimum  air  supply  may  cause  dense 
clouds  of  smoke  and  still  give  a  higher  evaporation  than  one  made 
smokeless  by  a  very  large  excess  of  air.     There  will  be  some  loss  due 
to  carbon  monoxide  and  unburned  carbon  or  soot  in  the  former  case, 
but  this  may  be  more  than  offset  by  the  excessive  losses  caused  by  the 

*  See  footnote,  p.  88.     f  See  footnote,  p.  37. 


FUELS  AND  COMBUSTION 


39 


heat  carried  away  in  the  chimney  gases  in  the  latter.  Judging  from  the 
results  of  the  majority  of  steam  power  plants  using  bituminous  coal, 
even  those  recently  installed,  smokeless  and  efficient  combustion  is 
not  readily  effected  and  the  problem  is  far  from  being  satisfactorily 
solved. 

Smoke  has  become  such  a  public  nuisance,  particularly  in  the  larger 
cities,  that  special  ordinances  prohibiting  its  production  have  been 
enacted  and  violators  are  subject  to  heavy  fines.  Effective  enforce- 
ment of  these  ordinances  renders  smoke  production  very  costly  and 
the  problem  of  smokeless  combustion  becomes  a  momentous  one. 

The  subject  of  smoke  prevention  and  smoke-prevention  devices  is 
discussed  at  some  length  in  Chapter  V. 

30.  Radiation  and  Minor  Losses.  —  These  losses  are  usually  deter- 
mined by  difference.  That  is,  the  difference  between  the  heat  repre- 
sented in  the  steam  and  the  losses  just  mentioned  is  charged  to  radia- 
tion, leakage,  and  unaccounted  for.  Summing  up  the  various  losses 
we  have 


Excellent 

Good 

Average 

Poor 

Practice. 

Practice. 

Practice. 

Practice. 

Per  Cent. 

Per  Cent. 

Per  Cent. 

Per  Cent. 

Heat  giv.en  to  steam  

80 

70 

60 

50 

Loss  in  chimney  erases 

12 

18 

24 

30 

Loss  due  to  carbon  burning  to  CO    

0 

1 

2 

3 

Loss  of  fuel  through  grate                .    .  . 

0  5 

1 

2 

3 

Loss  due  to  moisture  in  coal,-  moisture 

in  air,  and  hydrogen  in  fuel  

3.0 

3 

3 

3.5 

Smoke,  soot,  etc         

0 

0  5 

1 

1  5 

Radiation  and  minor  losses  

4  5 

6  5 

g 

9 

31.  Size  of  Coal.  —  Bituminous.  Coal  is  usually  marketed  in 
different  sizes,  ranging  from  lump  coal  to  screenings.  The  latter  furnish 
by  far  the  greater  part  of  the  stoker  fuel  used.  For  maximum  efficiency 
coal  should  be  uniform  in  size.  With  hand-fired  furnaces  there  is 
usually  no  limit  to  its  fineness  and  larger  sizes  can  be  used  than  with 
stokers.  As  a  rule  the  percentage  of  ash  increases  as  the  size  of  coal 
decreases.  This  is  due  to  the  fact  that  all  of  the  fine  foreign  matter 
separated  from  larger  coal,  or  which  comes  from  roof  or  floor  of  the 
mine,  naturally  finds  its  way  into  the  smaller  coal.  The  size  best  adapted 
for  a  given  case  is  dependent  upon  the  intensity  of  draft,  kind  of  stoker 
or  grate,  and  the  method  of  firing,  and  its  proper  selection  often  affords 
an  opportunity  to  effect  considerable  economy.  Some  idea  of  the 
influence  of  the  size  of  screenings  on  the  capacity  and  efficiency  of  a 


40 


STEAM  POWER  PLANT  ENGINEERING 


boiler  in  a  specific  case  is  illustrated  in  Fig.  6.  The  curves  are  plotted 
from  a  series  of  tests  conducted  with  Illinois  screenings  on  a  500-horse- 
power  B.  &  W.  boiler,  equipped,  with  chain  grates,  at  the  power  house 
of  the  Chicago  Edison  Company. 

Influence  of  Thickness  of  Fire.  —  See  paragraph  76. 

Size  of  Coal:  Some  Characteristics  of  Coal  as  affecting  Performances  with  Steam 
Boilers :  Jour.  West.  Soc.  Engrs.,  Oct.,  1906,  p.  528.  Small  Size  Anthracite :  Eng.  and 
Min.  Jour.,  Dec.  22,  1904.  The  Economy  of  Small  Size  Coals  for  Power  Plants  :  Eng. 
Mag.,  Feb.,  1905. 


40 


2SSS1 


\ 


Influence  of  Size  of  Coal  on  the  Capacity 
and  Efficiency  of  a  B.&W.Boiler,  Chain  Grate 
Heating  Surface  5000  Sq.Ft. 
Superheating  Surface  1000  Sq.Ft 


V 


\ 


\\ 


1000 


800 


400 


200  20  I 


1.25  1.00  0.75  0.50  0.25 

Size  of  Coal  in  Inches 

FIG.  6.   Influence  of  Size  of  Coal  on  Boiler  Capacity  and  Efficiency. 


32.  Washed  Coal.  —  The  washing  of  coal  is  for  the  purpose  of 
separating  from  it  such  impurities  as  slate,  sulphur,  bone  coal,  and  ash. 
All  of  these  impurities  show  themselves  in  the  ash  when  the  coal  is 
burned.  Screenings  contain  anywhere  from  5  per  cent  to  25  per  cent 


FUELS  AND  COMBUSTION 


41 


of  ash  and  from  1  per  cent  to  4  per  cent  of  sulphur.  Washing  eliminates 
about  50  per  cent  of  the  ash  and  some  of  the  sulphur.  Table  6  gives 
some  idea  of  the  effects  of  washing  upon  a  number  of  grades  of  coal. 
The  evaporative  power  of  the  combustible  is  practically  unaffected 
by  washing  and  the  greater  part  of  the  water  taken  up  by  the  coal  is 
removed  by  thorough  drainage.  Many  coals  otherwise  worthless  as 
steam  coals  are  rendered  marketable  by  washing.  Washed  coals  are 
usually  graded  as  follows : 


Size. 

Screens. 

No.  1 

Over  If 

Under  2* 

2 

1J 

1^ 

3 

I 

11 

4 

i 

5 

; 

Numbers  3  and  4  are  excellent  sizes  for  use  in  connection  with  stokers 
and  No.  5  is  well  adapted  for  hand  furnaces  where  smoke  prevention  is 
essential. 

TABLE  6. 
EFFECT    OF  WASHING    ON    BITUMINOUS    COALS. 

(Journal  W.S.E.,  December,  1901.) 


Before  Washing. 
(Per  Cent.) 

After  Washing. 
(Per  Cent.) 

Ash. 

Sul- 
phur. 

Fixed 
Carbon. 

Ash. 

Sul- 
phur. 

Fixed 
Carbon. 

Belt  Mountain,  Mont  
Wellington  Colliery  Co.,  Van- 
couver Island  (new  coal)  .... 
Alexandria  Coal  Co.,  Crabtree, 
Pa 

18.74 
35.00 

10.60 
18.00 

16.30 
15.80 

25.30 
13.77 
9.48 

3.34 

1.30 

0.57 
1.90 

1.05 
0.78 

43.72 
38.00 

44.00 
45.90 

5.56 
8.90 

6.21 
4.20 

9.70 
8.00 

8.50 
4.30 
4.85 

2.40 

0.61 

0.40 
0.87 

0.89 
0.69 

48.39 
56.90 

DeSoto,  111  
Northwestern      Improvement 
Co  ,  Roslyn,  Wash 

57.00 

47.86 
50.90 

47.20 
54.82 
63.00 

Luhrig  Coal  Co.,  Zaleski,  Ohio 
Rocky   Ford    Coal    Co.,    Red 
Lodge,  Mont  

37.80 
49.04 
55.00 

Buckeye  Coal   and   Ry.   Co., 
Nelsonville,  Ohio  

New  Ohio  Washed  Coal  Co., 
Carterville,  111 

42  STEAM  POWER  PLANT  ENGINEERING 

Modern  Method  of  Coal  Washing :  Eng.  and  Min.  Jour.,  May  9,  1903,  Oct.  13, 1904. 
Principles  of  Coal  Washing :  Mines  and  Min.,  Aug.,  1903.  Bituminous  Coal  Washing  : 
Mines  and  Min.,  April,  1905.  Washing  of  Bituminous  Coals  by  Luhrig  Process :  Jour. 
West.  Soc.  Engrs.,  Dec.,  1901.  Coal  Washing :  Jour.  Soc.  Chem.  Ind.,  April  30,  1904. 
Anthracite  Washeries :  Eng.  and  Min.  Jour.,  April  28,  1906;  Col.  Guard,  April 
20,  1906;  Trans.  Am.  Inst.  Min.  Engrs.,  Nov.,  1905.  Studies  on  Coal  Wash- 
ing :  Col.  Guard,  Nov.  21,  1902.  Coal  Washing  by  Stuart  System :  Mines  and  Min., 
Dec.,  1903.  Coal  Washing  at  Collinsville,  III:  Mines  and  Min.,  Sept.,  1901.  Bellevue 
Washery  of  D.  L.  and  W.  R.R.  Co.,  Scranton,  Pa.:  Mines  and  Min.,  June,  1903.  Coal 
Washery  at  Howe,  Indian  Territory  :  Mines  and  Min.,  March  14,  1904.  Eastern  Coal 
and  Coke  Co.'s  Washery  at  Kansas :  Eng.  and  Min.  Jour.,  Sept.  20,  1902. 

33.  Purchasing  Coal.* —  Engineers  fail  to  agree  as  to  the  specifica- 
tions best  suited  for  the  purchase  of  coal.  Some  extensive  purchasers 
require  elaborate  analyses  and  others  specify  only  the  size  and  grade 
of  the  fuel.  Every  essential  requirement  of  the  purchaser  may  be 
fulfilled  by  confining  them  to  the  four  following  characteristics: 

Moisture. 

Ash. 

Size  of  coal. 

Calorific  value  of  the  coal. 

Although  moisture  is  a  great  and  uncertain  variable,  and  the  producer 
can  exercise  no  control  over  this  factor,  still  the  purchaser  should  pro- 
tect himself  against  excessive  moisture  by  stipulating  an  amount  con- 
sistent with  the  average  inherent  moisture  in  the  coal,  and  proper  penalty 
should  be  fixed  for  delivery  in  excess  of  the  amount  allowed,  a  corre- 
sponding bonus  being  paid  for  delivery  of  less  than  contract  amount. 
Considerable  attention  should  be  given  to  the  percentage  of  earthy 
matter  contained.  The  amount  of  earthy  matter  usually  fixes  the 
heating  value  of  the  coal,  since  the  heating  value  of  the  combustible 
is  practically  constant.  The  effect  of  ash  on  the  heat  value  of  Illinois 
screenings  as  fired  under  a  B.  &  W.  boiler  with  chain  grate  is  shown 
in  Fig.  7.  This  value  varies  with  the  different  types  of  boilers,  grates, 
and  furnaces,  but  is  substantially  as  illustrated.  The  amount  of  refuse 
in  the  ash  pit  is  always  in  excess  of  the  earthy  matter  as  reported  by 
analysis. 

The  maximum  allowable  amount  of  sulphur  is  sometimes  specified, 
since  some  grades  of  coal  high  in  sulphur  cause  considerable  clinker- 
ing.  But  sulphur  is  not  always  an  indication  of  a  clinker-producing 
ash,  and  a  more  rational  procedure  would  be  to  classify  a  coal  as 
clinkering  or  non-clinkering  according  to  its  behavior  in  the  particular 
furnace  in  question,  irrespective  of  the  amount  of  sulphur  present. 
An  analysis  of  the  various  constituents  of  the  ash  is  necessary  to 

*  See  also  Selection  of  Coal  for  Boiler  Furnaces,  by  D.  T.  Randall,  Power  &  Engr., 
Apr.  6,  1909,  p.  642. 


FUELS  AND  COMBUSTION 


43 


determine  whether  or  not  the  sulphur  unites  with  them  to  produce  a 
fusible  slag,  and  as  such  analyses  are  usually  out  of  the  question  on 
account  of  the  expense  attached,  they  may  well  be  omitted. 

The  heating  value  of  the  coal  as  determined  by  a  sample  burned  in  an 
atmosphere  of  oxygen  does  not  give  its  evaporative  power,  since  this 


100 
90 
80 

ro 

60 
50 
40 
30 
<20 
10 

\ 

\ 

\ 

\ 

X 

\ 

X. 

\ 

\ 

\ 

\ 

\ 

V 

\ 

\ 

\ 

\ 

\ 

\ 

\ 

\ 

\ 

Influence  of  Ash  on  Fuel  Value  of  Dry 
Coal.    (Illinois  Screenings) 
B.&  W.  Boiler,  Chain  Grate. 
Screenings  with  12.5  Per  Cent  Ash 
taken  at  100. 

\ 

\ 

\ 

'our.J 

.W.E 

.Oct.l 

W6P. 

i 

10 


20  30 

Per  Cent  of  Ash  in  Dry  Coal. 


FIG.  7.   Influence  of  Ash  on  Fuel  Value  of  Dry  Coal. 

depends  largely  upon  the  composition  of  the  fuel,  character  of  grate, 
and  conditions  of  operation.  It  merely  serves  as  a  basis  upon  which  to 
determine  the  efficiency  of  the  furnace.  In  large  plants  where  a  number 
of  grades  of  fuel  are  available  it  is  customary  to  conduct  a  series  of 
tests  with  the  different  grades  and  sizes,  and  the  one  which  evaporates 


44  STEAM  POWER  PLANT  ENGINEERING 

the  most  water  for  a  given  sum  of  money,  other  conditions  permitting, 
is  the  one  usually  contracted  for.  In  designing  a  new  plant  particular 
attention  should  be  paid  to  the  performance  of  similar  plants  already 
in  operation,  and  that  fuel  and  stoker  should  be  selected  which  are  found 
to  give  the  best  returns  for  the  money.  Where  smoke  prevention  is  a 
necessity  the  smoke  factor  greatly  influences  the  choice  of  fuel  and 
stoker. 

See  paragraph  416. 

Testing  and  Purchasing  Coal  for  Steam  Plants  :  Eng.  News,  Feb.  7,  1907;  Eng.  Rec., 
Sept.  22,  1906,  p.  326;  Engr.  U.S.,  Aug.  15,  1907;  Bulletin  No.  339  U.S.  Geological 
Survey,  1908.  Coal  for  Hand-Fired  Furnaces:  Nat.  Engr.,  July,  1909. 

34.  Powdered  Coal.  — The  value  of  powdered  coal  as  a  fuel  for  steam 
boiler  plants  has  long  been  known,  and  appliances  for  pulverizing  and 
feeding  the  coal  have  been  on  the  market  for  a  number  of  years.  How- 
ever, despite  the  many  advantages  of  powdered  fuel  and  the  apparent 
success  of  some  of  the  systems  of  burning  it,  little  progress  has  been 
made  toward  its  general  adoption. 

Some  of  the  advantages  obtained  in  burning  powdered  coal  are: 

a.  Complete  combustion  and   total   absence   of  smoke.     The   coal 
in  the  form  of  dry  impalpable  dust  is  induced  or  forced  into  the  zone 
of  combustion,  where  each  minute  particle  is  brought  into   contact 
with  the  necessary  amount  of  air  and  complete  oxidation  is  effected 
without  the  excess  of  air  which  accompanies  the  firing  with  lump  coal, 
provided   the   furnace   is   properly    proportioned.     With    a    properly 
designed  setting  there  is  complete  absence  of  smoke. 

b.  A  cheaper  grade  of  bituminous  coal  may  be   burned,  since   the 
per  cent  of  ash  and  moisture  has  little  effect  on  the  completeness  of 
combustion  and  the  full  value  of  the  fuel  is  more  nearly  realized  than 
with  ordinary  firing. 

c.  The  plant  may  be  rapidly  forced  above  its  rated  capacity  and 
sudden  demands  for  power  readily  met. 

d.  The  labor  of  firing  is  reduced  to  a  minimum.    . 

Pulverized  Fuel :  Eng.  Mag.,  Jan.,  1908;  Jour.  West.  Soc.  Engrs.,  Feb.,  1904;  Am. 
Elecn.,  Sept.,  1901.  Coal  Dust  far  Steam  Making :  Engr.  U.S.,  Feb.  15, 1899.  Burn- 
ing Pulverized  Coal:  Eng.  and  Min.  Jour.,  Dec.  31,  1903,  May  12,  1906;  Jour.  Assn. 
Eng.  Soc.,  July,  1903.  Use  of  Pulverized  Coal  under  Steam  Boilers :  Eng.  News, 
April  1,  1904;  Power,  March,  1904,  April,  1904.  Coal  Dust  Firing:  Eng.  and  Min. 
Jour.,  Dec.  16,  1905.  Coal  Dust  Fuel:  Engr.,  Lond.,  Jan.  31,  1896;  Engr.  U.S., 
April  1,  1903;  Eng.  News,  Feb.  20,  1902. 

35.  Depreciation  of  Powdered-Coal  Furnaces.  —  To  withstand  the 
intense  heat  of  combustion,  brickwork  of  the  highest  quality  is  essential, 
since  common  fire  brick  are  soon  reduced  to  a  liquid  slag.  A  good  quality 


FUELS  AND  COMBUSTION  45 

of  fire  brick  will  withstand  the  heat  for  several  months  without  renewals 
provided  the  furnace  is  properly  enclosed,  otherwise  the  strain  of 
expansion  and  contraction  due  to  alternate  heating  and  cooling  will 
crack  the  brick.  Excellent  results  have  been  obtained  from  the  use  of 
bricks  composed  chiefly  of  the  refuse  from  a  carborundum  slag,  but  the 
high  cost  has  prevented  their  general  use. 

36.  Storing  Powdered  Fuel.  —  Most  cities  limit  the  storage  of  pow- 
dered coal  to  such  a  small  quantity  as  to  prohibit  the  use  of  fuel  feeders 
of  the  "  dust  feed  "  type  in  plants  of  any  size  not  provided  with  a 
pulverizing  and  crushing  system.     Coal  dust  mixed  with  air  is  often 
claimed  to  be  of  an  explosive  nature  and  many  accidents  are  reported 
to  have  resulted  from  this  cause.     Many  engineers,  however,  refute 
this  on  the  basis  of  experiments  which  show  that  explosion  can  only 
occur  at  temperatures  high  enough  to  drive  off  the  volatile  gases.* 

37.  Rate  of  Combustion  with  Powdered  Fuel.  —  In  forcing  large  quan- 
tities of  dust  into  the  furnace  the  velocity  imparted  to  the  particles  may 
be  so  great  as  to  carry  them  beyond  the  zone  of  combustion  before  oxida- 
tion is  complete,  with  the  result  that  the  flues,  and  the  back  of  the  fur- 
nace, become  covered  with  unconsumed  carbon.    So  much  depends  upon 
the  depth  of  the  furnace  and  the  arrangement  of  the  regenerative  surface 
that  no  specific  figures  can  be  given  as  to  the  maximum  rate  of  combus- 
tion that  can  be  efficiently  effected.    At  ordinary  rates  of  combustion  the 
small  particles  of  fuel  are  completely  oxidized  while  in  the  combustion 
chamber  and  there  is  total  absence  of  smoke.     The  use  of  anthracite 
coal  is  practically  excluded  from  this  type  of  stoker  unless  mixed  with 
coal  high  in  volatile  matter.     This  is  due  to  the  fact  that  fixed  carbon 
burns  more  slowly  than  the  hydrocarbon  gases  and  the  temperature  of 
ignition  is  higher,  hence  the  most   gentle  draft   will  carry  away  the 
particles  before  they  are  completely  consumed.     With  fuels  high  in  vola- 
tile matter  the  hydrocarbons  are  distilled  at  a  comparatively  low  tem- 
perature, forming  an  inflammable  gas  which  burns  rapidly  with  the  fixed 
carbon.     A  mixture  of  30  per  cent  bituminous  and  70  per  cent  anthra- 
cite has  been  successfully  burned  in  the  powdered  form. 

38.  Cost  of  Pulverizing  Coal.  —  In  stokers  of  the  "  Aero  Pulverizer  " 
type  in  which  the  grinding  and  feeding  are  carried  on  simultaneously 
in  a  self-contained  apparatus,  the  power  consumed  varies  from  2  per 
cent  to  10  per  cent  of  the  total  power  developed,  depending  upon  the 
nature  of  the  fuel,  the  load  factor,  the  efficiency  of  the  driving  mechan- 
ism, and  the  degree  of  fineness  of  the  powdered  fuel;  5  per  cent  is  a 
fair  average.     The  best  results  are  obtained  when  95  per  cent  of  the 
dust  will  pass  a  100  mesh  and  75  per   cent  a  200  mesh,  though  satis- 
factory results  have  been  obtained  with 'as  low  as  40  mesh.     Powdered 

*  See  Fuel,  Jan.  12,  1909,  p.  294. 


46 


STEAM  POWER  PLANT  ENGINEERING 


coal  in  the  open  market  ranges  from  25  cents  to  50  cents  a  ton  above 
the  price  of  the  same  coal  in  the  form  of  screenings. 

39.  Efficiency  of  Powdered-Coal  Furnaces.  —  Table  7  gives  the 
results  of  a  comparative  test  of  a  140-horse-power  Babcock  &  Wilcox 
boiler,  hand  fired,  vs.  coal-dust  feeder.  The  test  was  conducted  by. 
the  engineering  staff  of  the  McCormick  Harvester  Company  at  Chicago, 
Illinois,  and  the  results  were  obtained  with  boilers  working  under  normal 
conditions.  The  dust  apparatus  was  a  modified  "  Ruhl  "  feeder,  and 
was  installed  by  the  C.  0.  Bartlett  &  Snow  Company  of  Cleveland,  Ohio. 
In  this  particular  test  the  efficiency  of  the  boiler  is  very  low  for  both 
hand  fired  and  dust-feed,  but  the  dust-feed  test  shows  an  efficiency  of 
10  per  cent  greater  than  that  of  the  hand  fired. 


TABLE   7. 

COMPARATIVE   TEST   OF    140-HORSE-POWER   BABCOCK   &  WILCOX   BOILER. 
Hand  Fired  vs.  Pulverized  Fuel. 


Boilers  fired  by 

machine 

hand 

Date                        

2-24-04 

2-3-04 

Duration  of  test  

8  hours 

8  hours 

Total  water  evap.  into  dry  steam  from 
and  at  212  degrees  

70,070 

45,673 

Average  gauge  pressure  
Average  feed-water  temperature,  Fahr.  .  .  . 
Average  stack-gas  temperature,  Fahr  
Kind  of  coal  used  

79.8  Ib, 
169.7 
506 
Westville,    Indi- 

79.41b. 
172.1 
458 
Westville 

Cost  of  coal  delivered  in  boiler  room  ready 
to  fire 

ana,  screenings 
pulverized    to 
40  mesh 

$2  10 

screenings. 

$1.72 

Total  weight  of  dry  coal  consumed  
Per  cent  of  ash  in  coal  determined  by  lab- 
oratory analysis  

9,373 
17.5 

8,413 
19.54 

Per  cent  of  ash  as  removed  from  ash  pit 
and  furnace 

none 

20.57 

Heating  value  of  coal 

12,555 

11,300 

Water  evap.  per  pound  of  fuel,  actual  con- 
ditions                     .   ...       

6.822  Ib. 

4.595  Ib. 

Equivalent  water  evap.  from  and  at  212 
degrees  per  pound  of  dry  fuel 

7  476  Ib. 

5.429  Ib. 

Equivalent  water  evap.  per  pound  com- 
bustible   

9.132  Ib. 

6.941  Ib. 

Horse  power  developed                 

254 

165.5 

Dry  fuel  per  hour  per  square  foot  grate 
surface                             .         

19.27 

18.65 

Equivalent    water   evap.    per   hour   per 

3.128 

2.039 

Cost  per  1,000  pounds  water  evaporated  (for 
fuel  ready  to  fire  only)                         .... 

$0.1455 

$0.177 

Efficiency  of  boiler  and  furnace  based  on 
coal  

55.5  per  cent 

41  .5  per  cent 

FUELS  AND  COMBUSTION  47 

A  comparison  of  a  number  of  tests  of  hand-fired  and  powdered-coal 
furnaces  with  different  types  of  feeders  shows  a  decided  gain  in  efficiency 
of  the  powdered  coal  over  the  hand-fired  where  the  fuel  is  of  a  low  grade. 
The  gain  becomes  less  marked  with  fuel  of  fair  quality  and  disappears 
entirely  with  good  fuel  and  properly  manipulated  automatic  stokers. 
A  test  made  by  G.  H.  Barrus  on  a  250-horse-power  B.  &.  W.  boiler  at 
the  General  Electric  Works  in  connection  with  a  coal-dust  feeder  manu- 
factured by  the  Phoanix  Investment  Company  of  New  York  gave  a  boiler 
and  furnace  efficiency  of  75.3  per  cent.  Subtracting  from  this  the  power 
consumption  of  5  per  cent  for  operating  the  crusher  and  feeder,  the  net 
efficiency  was  70.1  percent.  A  test  of  a  135-horse-power  return  tubular 
boiler  with  this  same  stoker  gave  a  combined  efficiency  of  boiler  and 
furnace  of  80  per  cent.  These  figures,  however,  have  been  equaled  and 
even  exceeded  in  special  hand-fired  automatic  stoker  tests,  and  only  a 
comparative  test  of  the  two  systems  under  similar  conditions  will  show 
their  respective  efficiencies. 

Tests  of  Pulverized  Fuel:  Engr.  U.S.,  April  1,  1904;  Engr.  Lond.,  Jan.  31,  May, 
1904;  Power,  May,  1904.  Comparative  Boiler  Tests  with  Ordinary  and  Pulverized 
Coal  Firing :  Eng.  Rec.,  March  12,  1904. 

40.  Furnaces  for  Burning  Powdered  Coal.  —  In  burning  ordinary  bulk 
coal  the  mass  of  incandescent  fuel  stores  up  a  sufficient  quantity  of  heat 
to  effect  the  distillation  and  ignition  of  the  volatile  matter  in  the  green 
fuel.     With  pulverized  coal  a  refractory  lining  is  necessary  to  bring 
about  the  same  result.     In  arranging  a  furnace  for  burning  powdered 
coal  in  connection  with  a  burner  of  the  forced  draft  type,  the  grate  bars 
are  removed,  ash  and  fire  doors  bricked  up,  and  the  nozzle  bricked  in 
tightly.     The  lower  surfaces  of  the  tubes  are  covered  and  the  whole 
forms   a  reverberatory  furnace.     With  the   natural   draft  system   of 
burner,  a  suitable  opening  is  left  in  the  brick  lining  of  the  ash  door  to 
allow  the  necessary  amount  of  air  for  combustion  to  enter.     Considerable 
difficulty  is  found  with  delivery  nozzles  in  the  formation  of  slag  in  the 
outlet  and  in  their  rapid  destruction  on  account  of  the  intense  heat.     A 
water- jacketed  cast-iron  nozzle  is  said  to  satisfactorily  overcome  these 
objections. 

41.  Draft  for  Powdered  Fuel.  —  A  study  of  a  number  of  tests  of 
boilers  burning  powdered  coal  shows  that  the  necessary  draft  is  very 
low  and  ranges  from  0.05  to  0.2  of  an  inch  of  water  and  averages  not  far 
from  0.1  inch. 

42.  Types  of  Powdered-Coal  Burners.  —  Powdered-coal  burners  may 
be  grouped  into  two  general  classes: 

1.  The  dust-feed  burner,  in  which  the  coal  is  supplied  in  the  powdered 
form,  and 


48 


STEAM  POWER  PLANT  ENGINEERING 


2.   The  self-contained  burner,  in  which  the  coal  is  crushed,  pulver- 
ized, and  fed  to  the  furnace  simultaneously. 
The  dust  may  be  fed  into  the  furnace  by 

1.  Natural  draft. 

2.  Mechanical  means,  or  by 

3.  Forced  draft. 

The  following  outline  gives  a  classification  of  a  few  of  the  best  known 
coal-dust  burners: 

Natural  Draft       fpinther 


Natural  Draft 


Forced  Draft  . 


Feed 
Brush  Feed 

Blower  Feed 


jWegener 
Schwartzkopff 

/Cyclone 
\Triumph 


Dust  Feed 


Compressed  Air   /Eng  and  Powdered 
(.Fuel  Company 


Paddle  Wheel 


(Ideal 
lAero-Pulverizer 


Self-contained 


43.   Pinther  Apparatus.  —  Fig.  8  shows  a  section  through  a  Pinther 
coal-dust  feeder,  illustrating  the  principles  of  the  "  natural  draft  feed  " 

type.  The  powdered  coal  is 
placed  into  hopper  B,  from 
which  it  is  fed  by  rollers  a,  a 
into  the  chamber  leading  to 
the  furnace  C.  The  dust  falls 
in  a  thin  stream  and  is  caught 
up  by  the  current  of  air  and 
drawn  into  the  furnace  as 
indicated.  The  furnace  is 
lined  with  refractory  material 
^ea^ec^  *°  a  sufficiently  high 
temperature  to  ignite  the  fuel 
and  burn  it  in  suspension.  The 
chief  drawback  to  a  burner 
of  this  type  is  its  limited 
capacity.  Any  attempt  to 
feed  large  quantities  of  fuel 
into  the  furnace  necessitates 
such  a  strong  current  of  air  as 
to  carry  the  particles  of  dust  beyond  the  zone  of  combustion  before  they 
are  completely  consumed.  Within  the  limits  of  its  capacity  it  is  an 
efficient  and  simple  apparatus,  but  is  open  to  the  same  objection  as  all 


FURNACE 


FIG.  8.    Pinther  Coal-dust  Feeder. 


FUELS  AND  COMBUSTION 


49 


FURNACE 


burners  of  this  type  in  that  it  necessitates  the  storage  of  powdered 
coal.     This  apparatus  is  not  much  in  evidence  in  boiler  plants. 

44.  Schwartzkopff  Apparatus.  —  Fig.  9  shows  a  section  through  a 
Schwartz  kopff  feeder,  illustrating 

the  principles  of  the  brush-feed, 
natural-draft  system.  It  is  a 
very  simple  and  practical  dust 
feeder,  though  open  to  the  objec- 
tion of  all  systems  which  require 
the  coal  to  be  ground  and  pulver- 
ized in  separate  machines.  The 
fuel  is  placed  in  a  hopper  and  its 
supply  to  the  brush  is  regulated 
by  the  hand  screw  A  and  the 
spring  plate  bottom  of  the 
hopper.  The  brush,  consisting 
of  a  number  of  flat  steel  leaves 
^j  inch  by  J  inch  wide,  revolves 
at  a  high  speed,  1000  to  1200 
r.p.m.  and  forces  the  dust  into 
the  furnace.  The  air  for  combus- 
tion is  admitted  either  through 
the  grates  in  the  ordinary  way  or 
through  the  lower  chamber  of 
the  burner.  To  prevent  the  dust 
from  bridging  in  the  hopper,  a  small  hammer  C  is  fitted  to  the  brush  so 
that  it  will  strike  the  plate  D  and  agitate  the  dust.  This  apparatus  is 
meeting  with  much  success  in  connection  with  annealing  furnaces,  but 
is  still  in  the  experimental  state  as  far  as  boiler  firing  is  concerned. 

45.  Aero-Pulverizer  Apparatus.  —  Fig.  10  gives  a  general  view  of  the 
Aero-Pulverizer  Company's  apparatus,  and  is  a  typical  example  of  a  self- 
contained  system.     It  is  very  compact,  occupying  a  floor  space  of  only 
30  by  77  inches,  and  is  capable  of  burning  300  to  1500  pounds  of  coal  per 
hour.     It  consists  essentially  of  four  interior  communicating  chambers 
of  successively  increased  diameter  in  which  paddles  revolve  on  arms 
with  corresponding  increased  radii.      The  largest  chamber  contains  a 
fan,  the  function  of  which  is  to  draw  the  pulverized  material  successively 
from  one  chamber  to  another  and  to  finally  deliver  it  through  the  exit 
in  the  fan  chamber  under  the  impetus  of  a  forced  draft.     There  are  two 
adjustable  inlets  for  air  at  the  feed  of  the  machine  through  which  is 
introduced  the  amount  of  air  required  for  pulverizing  purposes.     The 
apparatus  may  be  belt  driven  or  direct  connected  and  runs  at  about 


FIG.  9.    Schwartzkopff  Coal-dust  Feeder. 


50 


STEAM  POWER  PLANT  ENGINEERING 


1600  r.p.m.,  requiring  from  6  to  15  horse  power  for  its  operation.  It  is  a 
complete  dust  fuel  feeding  system  on  one  bed  plate  comprising  a  pul- 
verizer, fan,  coal  feeder,  hopper,  and  air  dampers.  The  operation  is  as 
follows:  Coal  previously  crushed  to  nut  size  is  fed  to  the  hopper  from 


TOP    CASING    THROWN    BACK 
FOR    INSPECTION 


COAL 
HOPPER 


AUTOMATIC 
FEEDER 


FIG.  10.    Aero-Pulverizer  Coal-dust  Feeder. 

the  bottom  of  which  it  is  transferred,  with  the  necessary  air  for  com- 
bustion, to  the  pulverizer  chamber.  The  coal,  passing  into  the  pul- 
verizer, is  thrown  out  radially  by  centrifugal  force,  due  to  the  rapidly 
revolving  arms  and  bats,  and  is  reduced  to  a  dust  by  percussion  and 
attrition.  The  dust  is  moved  over  the  ends  of  the  bats  and  into  the  fan 
chamber  from  which  it  is  blown  into  the  furnace.  This  apparatus  will 
successfully  pulverize  and  feed  coal  containing  as  much  as  10  per  cent 
moisture. 


RICTION    DISK 


FIG.  11.     Triumph  Coal-dust  Feeder. 

46.  Triumph  Apparatus.  —  Fig.  1 1  illustrates  the  Triumph  coal-dust 
feeder  as  designed  by  the  C.  O.  Bartlett  &  Snow  Company,  Cleve- 
land, Ohio. 


FUELS  AND  COMBUSTION  51 

The  .coal  is  fed  from  storage  bin  to  hopper  A  and  feed  worm  B. 
The  latter  forces  it  down  spout  F  directly  to  delivery  tube  D,  where  it 
is  caught  by  the  air  draft  and  fed  into  the  furnace. 

The  amount  of  feed  depends  upon  the  speed  of  the  feed  worm,  which 
is  driven  by  the  friction  disk  /  against  the  flange  plate  H.  This  disk  is 
moved  in  or  out  by  handle  so  as  to  get  any  speed  desired.  The  air  is 
furnished  by  fan  C,  the  amount  being  controlled  by  valve  E. 

DESCRIPTION  OF  COAL-DUST  BURNERS. 

Aero-Pulverizer  System:  Eng.  News,  Nov.  28,  1901,  p.  415;  Eng.  Rec.,  May  25, 

1901,  p.  506;  Power,  March,  1904. 

Cyclone  System:  Engr.  U.S.,  April  1,  1903,  p.  272;  Eng.  News,  Nov.  28,  1901, 
p.  415;  Power,  March,  1904. 

Davis  Pulverizer:  Jour.  Asso.  Eng.  Soc.,  July,  1903;  Engr.  U.S.,  April  1,  1903. 

Ideal:  Am.  Elecn.,  April,  1902,  p.  196;  Power,  March,  1904. 

Miscellaneous  Coal  Dust  Burners:  Am.  Elecn.,  Sept.,  1901,  p.  434;  Engr.,  Lond., 
Sept.  11,  1896;  Engng.,  Jan.  15,  1897;  Power,  Aug.,  1903;  St.  Ry.  Rev.,  Vol.  8-187, 
1898;  Engr.  U.S.,  April  1,  1904. 

Rowe :  Engr.  U.S.,  Jan.  1,  1903,  p.  93;  April  1,  1903,  p.  272;  Eng.  News,  Dec.  25, 

1902,  p.  548;  Eng.  Rec.,  Dec.  20,  1902,  p.  591. 

Schwartzkopff :  Am.  Elecn.,  Jan.,  1902;  Eng.  News,  Feb.  20, 1902;  Power,  March,. 
1904. 

Wegener  :  Cassier's,  March,  1896,  p.  501;  Power,  March,  1904;  Eng.  Mag.,  March, 
1896,  p.  1158,  Aug.,  1896,  p.  964,  Oct.,  1898,  p.  125;  Eng.  News,  Sept.  16,  1897, 
p.  189, 

47.  Fuel  Oil.  —  The  recent  development  of  oil  wells  in  the  Western 
and  Gulf  States,  with  the  consequent  enormous  increase  in  production, 
has  given  a  marked  impulse  to  the  use  of  crude  oil  for  fuel  purposes  in 
steam  power  plants.  Where  economic  and  commercial  conditions 
permit,  it  is  the  most  desirable  substitute  for  coal.  The  total  absence 
of  smoke  and  ashes,  prompt  kindling  and  extinguishing  of  fires,  extreme 
rate  of  combustion,  and  ease  with  which  it  can  be  handled  and  con- 
trolled are  marked  advantages  in  favor  of  fuel  oil.  The  reduction  in 
volume  and  weight  over  an  equivalent  quantity  of  coal  for  equal  heat- 
ing values  and  the  increase  in  boiler  efficiency  are  factors  of  no  mean 
importance,  particularly  in  connection  with  marine  or.locomotive  work. 
In  stationary  work  the  chief  objections  are  the  difficulty  in  securing 
ample  storage  capacity  and  the  increased  rate  of  insurance.  An  objec- 
tion sometimes  raised  against  oil  fuel  is  the  increased  depreciation  of 
the  setting,  but  in  a  well-designed  setting  this  figure  is  only  nominal  and 
of  secondary  importance.  However,  in  spite  of  the  many  advantages 
presented  in  the  use  of  fuel  oil  for  power  plant  purposes,  the  limited 
supply  and  constant  fluctuation  in  price  prevent  its  adoption  as  a 
general  fuel,  and  limit  its  use  to  the  plants  most  favorably  located. 


52 


STEAM  POWER  PLANT  ENGINEERING 


Crude  Oil  Burning :  Power,  March,  1907;  Engr.  U.S.,  Dec.  15, 1905,  March  1,  1906; 
Am.  Elecn.,  Aug.,  1903,  p.  396;  Engng.,  March  28,  1902,  p.  140;  Eng.  Mag.,  May, 
July,  Sept.,  1903;  Eng.  News,  June  19,  1902,  p.  501;  Cassier's,  May,  1901,  p.  61; 
Engr.,  Lond.,  Dec.  9,  1904;  St.  Ry.  Jour.,  May  10,  1902,  p.  588;  Am.  Gas  Light 
Jour.,  May  12,  1902,  p.  695. 

TABLE   8. 
ANALYSES    OF   TYPICAL   AMERICAN    FUEL    OILS. 


Location. 

Authority. 

Physical  Properties. 

Specific 
Gravity. 
60°-70°  F. 

Flash 
Point. 
Deg.  F. 

Burn- 
ing 
Point. 
Deg.  F. 

Specific 
Viscosity. 

60°  F. 

185°  F. 

California  —  Crude  
Do  

Ed.  O'Neill.. 
....do  

0.9533 
0.9572 
0.7825 
0.9670 
0.866 

299.6 
373 
1.17 

4.7 

62 

196 
52 

64^5 
221 

77 

Do 

do 

Do 

do 

Kansas  —  Crude  
Louisiana  —  Crude  
Ohio  —  Distillate 

B.  F.  McFarland. 
C   E  Coates   . 

Deville     

0.887 
0.838 
0.826 
0.886 
0.841 

Do  

N.  W.  Lord  
Deville  

177 

212 

Pennsylvania  —  Crude  .  . 
Pennsylvania  —  Distillate 
W    Virginia  —  Crude 

....do  
do 

Wyoming  —  Crude  
Texas  —  Crude 

Colburn  

Denton  

0.92 
0.926 

142 
216 

181 
240 

Texas  —  Distillate  

U.  S.  Naval  Re- 

port 

Location. 

Authority. 

Chemical  Properties. 

C 

H 

O+N 

S 

B.T.U. 
per  Lb. 

California  —  Crude  
Do  

Ed  O'Neill  .    . 

85.75 
86.3 

11.3 
10.7 



0.668 
0.8 

18,797 
18,646 

do  

Do  

.....do  

Do  

....do  

Kansas  —  Crude           . 

B.  F.  McFarland. 
C  E  Coates. 

85.4 

13.07 

6.34 

19,814' 
18,718 
19,880 
17,930 
19,210 
18,400 
19,590 
19,060 
19,481 

Louisiana  —  Crude 

Ohio  —  Distillate       .    .  . 

Deville  
N.W.  Lord  

84.2 

13.1 

2.7 

Do   

Pennsylvania  —  Crude  .  . 
Pennsylvania  —  Distillate 
W.  Virginia  —  Crude  .... 
Wyoming  —  Crude  
Texas  —  Crude 

Deville  
....do  
....do  
Colburn 

82 
84.9 
84.3 

14.8 
13.7 
14.1 

3.2 
1.4 
1.6 

Denton 

84.6 
83.26 

10.9 
12.41 

2.87 
3.83 

1.63 
0.50 

Texas  —  Distillate  

U.  S.  Naval  Re- 
port 

48.  Chemical  and  Physical  Properties  of  Fuel  Oil.  —  From  Table  8 
it  will  be  seen  that  the  physical  properties  of  oils  from  different  localities 
in  the  United  States  differ  widely,  while  the  chemical  constituents  vary 


FUELS  AND  COMBUSTION 


53 


but  slightly.  For  example,  the  oils  given  in  the  table  differ  greatly  in 
volatility,  specific  gravity,  and  viscosity,  but  have  nearly  a  constant 
ratio  of  carbon  and  hydrogen  and  consequently  vary  but  slightly  in 
heating  value. 

A  good  deal  of  the  oil  produced  is  unfit  for  fuel  purposes  unless 
refined.  The  chief  impurities  are  sulphur,  earthy  matter,  and  water. 
Besides  interfering  with  the  free  burning  of  the  oil,  moisture  and  sulphur 
have  a  deleterious  effect  upon  the  boiler  and  furnaces,  and  should  not 
be  present  in  large  quantities.  Where  the  percentage  of  sulphur  is 
greater  than  4  per  cent,  the  depreciation  of  the  boiler  and  furnace 
offsets  the  gain  in  using  the  lower  grade  fuel.  Many  successful  processes 
of  removing  the  water  and  sulphur  are  on  the  market,  and  consequently 
crude  oil  high  in  sulphur  should  not  be  used  unless  the  depreciation 
element  has  been  taken  into  consideration. 

Oil  that  is  to  be  transported  or  stored  or  used  for  fuel  inside  of 
buildings  should  be  of  the  "  reduced  "  variety,  from  which  the  naphtha 
and  higher  illuminating  products  have  been  distilled.  The  gravities 
of  such  distillates  vary  from  20  to  25  degrees  Baume,  or  close  to  0.9 
specific  gravity,  and  their  flash  points  range  from  240  degrees  F.  to  270 
degrees  F.  This  variation  in  volatility  has  little  effect  on  the  heat 
value  of  the  oil,  since  the  ratio  of  carbon  to  hydrogen  varies  but  slightly 
in  the  various  distillates. 

One  barrel  of  oil  contains  42  gallons  and  weighs  from  310  to  332 
pounds  according  to  the  specific  gravity.  Compared  with  coal,  oil 
occupies  about  50  per  cent  less  space  and  is  35  per  cent  less  in  weight 
for  equal  heat  values.  The  comparative  heat  values  of  coal  and  oil 
are  approximately  as  follows: 


B.T.U.  per  Pound 
of  Coal. 

Pounds  of  Coal  Equal 
to  1  Barrel  of  Oil. 

Barrels  of  Oil  Equal 
to  1  Short  Ton  of 

Coal. 

10,000 

620 

3.23 

11,000 

564 

3.55 

12,000 

517 

3.87 

13,000 

477 

4.19 

14,000 

443 

4.52 

15,000 

413 

4.84 

Technical  Aspects  of  Oil  as  Fuel :  Junge,  Power,  Oct.,  1907,  p.  665. 

Petroleum  Oil  Fields :  Jour.  Soc.  Chem.  Ind.,  Oct.  15,  1902,  p.  1228.  Investiga- 
tion on  American  Petroleum :  Am.  Chem.  Jour.,  March,  1896,  p.  215.  The  Constit- 
uents of  Pennsylvania,  Ohio,  and  Canada  Petroleums:  Amer.  Chem.  Jour.,  Vol.  19-419. 
Composition  of  California  Petroleum  :  Amer.  Chem.  Jour.,  Vol.  19-796,  Vol.  25-253. 
Composition  of  Petroleum:  Amer.  Chem.  Jour.,  Vol.  28-165,  33-251.  Composition 


54 


STEAM  POWER  PLANT  ENGINEERING 


of  Texas  Petroleum:  Jour.  Amer.  Chem.  Soc.,  Feb.  9,  1901,  p.  264;  Soc.  Chem.  Ind., 
19-121,  20-237,  690.  Flashing  Points  of  Petroleum  :  Jour.  Soc.  Chem.  Ind.,  15-341. 
Origin  of  Petroleum :  Jour.  Soc.  Chem.  Ind.,  16-727,  1898.  Influence  of  Water  on 
Flash  Test  and  Combustion  Point  of  Petroleum  :  Chem.  News,  85-267.  The  Relation 
between  Some  Physical  Properties  of  Bitumens  and  Oils :  Eng.  Rec.,  Aug.  18,  1906. 

49.  Efficiency  of  Boilers  with  Fuel  Oil.  —  From  Table  14,  it  will  be 
seen  that  70  per  cent  is  a  high  figure  for  boiler  efficiency  in  regular 
service  when  good  coal  is  burned,  and  65  per  cent  a  fair  average.  With 
liquid  fuel  an  average  efficiency  of  4  to  6  per  cent  above  this  is  readily 
attained.  (See  Table  9.)  This  increase  in  efficiency  is  partly  due  to 
the  fact  that  the  oil  is  readily  broken  up  and  brought  into  intimate 
contact  with  the  necessary  air  for  combustion  and  the  loss  due  to  excess 
of  air  is  correspondingly  reduced.  The  results  of  tests  made  by  the 
Liquid  Fuel  Board  of  the  U.S.  Navy  show  that  oil  has  an  efficiency  of 
5  per  cent  greater  than  coal  for  the  same  rate  of  evaporation,  and  that 
the  boiler  capacity  may  be  increased  50  per  cent  above  that  possible 
when  burning  coal  and  still  maintain  the  same  efficiency.  The  max- 
imum efficiency  with  oil  was  attained  at  a  higher  rate  of  evaporation 
than  was  the  maximum  efficiency  when  coal  was  burned. 


TABLE  9. 

BOILER    EFFICIENCIES,    OIL    FUEL. 


H 

1 

Authority. 

Reference. 

Quality  of  Oil. 

Evapora- 
tion from 
and  at 
212°  F. 
Pounds. 

Efficiency 
of  Boiler 
and  Fur- 
nace. 
Per  Cent. 

1 

2 

Pacific  Light  and 
Power  Co.,  Los 
Angeles,  Cal. 

Eng.  Record,  Aug. 
6,  1904. 

California  Crude, 
18,607B.T.U.per 
pound. 
California  Crude, 
18,760  B.T.U. 

16.02 
15.66 

83.06 
80.6 

3 

4 

U.S.  Naval  Board 

1902    Report    of 
U.S.    Naval 
Liquid     Fuel 
Board. 

Reduced     Beau- 
mont,   19,480 
B.T.U. 
20,000  B.T.U  

14.43 
16.9 

71.5 
77.8 

5 

6 

Prof.  Williston.  .  . 

Engineering  Mag- 
azine, July,  1903. 

West  Va.  Crude, 
20,960  B.T.U. 
Texas  Crude, 
18,850  B.T.U. 

16.5 
15.9 

76 

76.8 

7 

Prof.  Denton  .... 

Power,  Feb.,  1902. 

Beaumont  Texas, 
19,060  B.T.U. 

'  15.5 

78.5 

8 

Wallsend  

Engineering,  Nov. 
6,  1902. 

Not  stated  

14.45 



FUELS  AND  COMBUSTION  55 

50.  Comparative  Evaporative  Economy  of  Oil  and  Coal.  —  In  deter- 
mining the  comparative  economy  of  coal  and  oil,  the  fixed  and  operating 
charges  must  be  considered  in  addition  to  the  cost  and  efficiency  of 
the  fuel.     From  the  market  quotation  on  oil  and  coal  and  the  com- 
parative heating  values  of  each  the  actual  cost  per  B.T.U.  is  readily 
obtained,  and  by  combining  this  with  the  relative  efficiencies  from  the 
furnace  standpoint  the  net  cost  of  the  fuel  is  obtained.     The  fixed 
charges  vary  with  the  location  and  size  of  the  plant  and  are  approxi- 
mately the  same  per  boiler  horse  power  for  a  given  location  in  both 
cases.     The  insurance  rates  may  be  greater  with  the  oil  fuel  and  the 
depreciation  of  the  boiler  setting  may  be  somewhat  larger,  but  in  a  well- 
constructed  furnace  the  latter  item  should  be  the  same  in  both  instances 
for  average  rates  of  combustion.     The  operating  charges  are  decidedly 
in  favor  of  the  oil  fuel,  since  no  ash  handling  is  necessary.     Oil  fuel  is 
readily  fed  to  the  furnace,  and  the  cost  of  attendance  may  be  materially 
less  than  with  coal  firing,  and  one  man  may  safely  control  from  eight  to 
ten  boilers.     Table  106,  Chapter  XVII,  gives  data  relative  to  the  cost 
of  producing  electrical  power  in  connection  with  oil-fired  steam  plants. 

Tests  of  Crude  Oil  as  a  Fuel:  Cassier's,  May,  1901, p.  61;  Power,  Feb.,  1902,  p.  8; 
Eng.  U.S.,  Jan.  16,  1905,  p.  90,  Feb.  1,  1905;  Eng.  Rec.,  Dec.  20,  1902  ;Eng.  Mag., 
July,  1902,  p.  615;  Eng.  News,  July  11,  1901,  p.  23;  Eng.  Rec.,  Aug.  6,  1904,  p.  175; 
Oct.  29,  1904,  p.  502. 

51.  Oil  Burners.  —  The  function  of  the  burner  is  to  atomize  the  oil 
to  as  nearly  a  gaseous  state  as  possible. 

Classification  of  a  few  well-known  burners: 

Mechanical  Spray  : 

Korting. 
Vapor  or  Carburettor : 

Durr. 

Harvey. 

Spray  Burners  : 
Outside  Mixers. 

a.  Booth. 

b.  Warren. 

Inside  Mixers. 
a.  Hammel. 
6.  Kirkwood. 

c.  Branch. 

d.  Williams. 


56 


STEAM  POWER  PLANT  ENGINEERING 


Oil  burners  for  burning  liquid  fuel  may  be  divided  into  three  general 
classes: 

1.  Mechanical  spray,  in  which  the  oil  previously  heated  to  a  tem- 
perature of  about  150  degrees  F.  is  forced  under  pressure  through 
nozzles  so  designed  as  to  break  it  up  into  a  fine  spray.  The  Korting 

Liquid  Fuel  Burner,  Fig.  12,  is  an 
example  of  this  type.  In  this 

O )  IKn  design  a  central  spindle' spirally 

grooved,  imparts  a  rotary  motion 
to  the  oil  and  causes  it  to  fly  into  a 
spray  by  centrifugal  force  on  issuing 
from  the  nozzle.  The  particles  of 
oil  are  burned  in  the  furnace  when 
they  come  in  contact  with  the  neces- 
sary air  to  effect  combustion.  This 
type  of  burner  is  little  used  in  this 


OIL 

FIG.   12.     Kbrting  Fuel  Oil  Burner. 


country  in   connection  with  power  plant  work,  but  is  meeting  with 
much  success  on  the  continent. 

2.  Vapor  burners,  or  carburettors,  in  which  the  oil  is  volatilized  in 
a  heater  or  chamber  and  then  admitted  to  the  furnace,  are  seldom  used 
except  in  connection  with  refined  oils,  as  the  residuals  from  crude  oil 
are  vaporized  only  at  a  high  temperature.     The  Durr  and  Harvey 
gasifiers  are  the  best  known  of  this  type. 

3.  Spray  burners  are  by  far  the  most  common  in  use.     In  this  type 
the  oil  is  held  in  suspension  and  forced  into  the  furnace  by  means  of  a 
jet  of  steam  or  compressed  air.     Spray  burners  are  designed  either  as 
outside  mixers,  in  which  the  oil  and  atomizing  medium  meet  outside  the 
apparatus,  or  inside  mixers,  in  which  the  oil  and  atomizing  medium 
mingle  inside  the  apparatus. 


FIG.  13.     Booth  Fuel  Oil  Burner. 


The  Booth  burner,  Fig.  13,  illustrates  the  principles  of  the  "  outside 
mixer  "  type  of  apparatus  and  is  in  use  on  the  Santa  Fe  Railroad.  In 
this  type  the  oil  flows  through  a  thin  slit  and  falls  upon  a  jet  of  steam 


FUELS  AND  COMBUSTION 


57 


which  atomizes  it  and  forces  it  into  the  furnace.     A  feature  of  this 
apparatus  is  its  simplicity  of  construction  and  freedom  from  clogging. 

Fig.  14  illustrates  the  Hammel  burner  as  used  at  the  power  house  of 
the  Pacific  Light  and  Power  Company,  Los  Angeles,  Cal.     Oil  enters  the 


G.I.   H. 

FIG.  14.    Hammel  Fuel  Oil  Burner. 


burner  under  pressure  and  flows  through  opening  D  to  the  mouth  of  the 
burner  where  it  is  atomized  by  the  steam  jets  issuing  from  slots  G,  H, 


STEAM 


OIL 

FIG.  15.  Branch  Fuel  Oil  Burner. 

and  7.  The  oil  is  preheated  to  facilitate  its  flow  through  the  supply 
system.  Plates  K-K  are  removable  and  are  easily  replaced  when  worn 
out  or  burned.  The  Hammel  burner  belongs  to  the  "  inside  mixers." 


58 


STEAM  POWER  PLANT  ENGINEERING 


A  few  well-known  types  of  "  inside  mixers  "  are  illustrated  in  Figs. 
14  to  16.  The  operation  is  practically  the  same  in  all  of  them  and  they 
differ  only  in  mechanical  details. 


FIG.  16.    Kirkwood  Fuel  Oil  Burner. 

C 


FIG.  17.    Williams  Fuel  Oil  Burner. 


The  Williams  burner,  Fig.  17,  differs  somewhat  from  the  others  in 
that  the  air  supply  passes  through  the  burner  and  mingles  with  the  oil 
and  steam  before  entering  the  furnace. 


FUELS  AND  COMBUSTION 


59 


The  simplest  and  most  reliable  burners  are  of  the  Hammel  type  and 
are  much  in  evidence  in  the  Pacific  States. 

Notes  on  Oil  Burners  using  Compressed  Air :  Power,  Nov.,  1904.  Report  of  U.S. 
Naval  Liquid  Fuel  Board  :  Engr.  U.S.,  Dec.  1,  1904.  Oil  Burners :  Engng.,  April  15, 
1904;  Am.  Engr.  and  R.R.  Jour.,  Sept.,  1901. 


FIG.  18.   Warren  Fuel  Oil  Burner. 


52.  Furnaces  for  Burning  Oil  Fuel.  —  Fig.  19  shows  the  construction 
of  a  typical  oil-burning  furnace  as  applied  to  a  250-horse-power  B.  &  W. 
water  tube  boiler  in  the  power  plant  of  the  Union  Loop  Elevated  Station, 
Chicago.  For  the  successful  burning  of  oil  the  furnace  should  be  so 
constructed  that  oxidation  of  the  fuel  is  complete  before  it  reaches  the 
tubes.  This  is  effected  by  arranging  the  refractory  lining  to  form  a 
sort  of  reverberatory  furnace  in  which  the  atomized  oil  is  vaporized  and 
mixed  with  the  necessary  air  for  combustion.  The  air  is  preheated  in 
its  passage  beneath  the  lower  lining  of  the  furnace  and  the  supply  is 
regulated  by  a  suitable  damper.  The  regulation  of  air,  steam,  and  oil 
for  the  burner  is  a  very  delicate  operation  and  requires  considerable 
skill  for  efficient  results.*  In  the  particular  furnace  illustrated  in  Fig.  19 
the  flame  impinges  against  a  cellular  wall  of  fire  brick  before  it  reaches 

*  For  a  modern  and  highly  efficient  system  of  oil  fuel  feeding  and  regulation,  see 
Power,  Dec.  29,  1908,  p.  1108. 


60 


STEAM  POWER  PLANT  ENGINEERING 


the  bridge  wall.     The  bricks  are  loosely  stacked  and  are  readily  removed 
when  burned  out.     They  tend  to  save  the  lining  of  the  bridge  wall  arch 

and  insure  a  more  intimate 
mixture  of  air  and  oil  in  the 
combustion  chamber.  The 
supply  of  crude  oil  is  limited 
compared  with  that  of  coal 
and  the  price  is  subject  to 
sharp  fluctuations,  and  conse- 
quently the  cost  may  prove 
prohibitive  for  fuel  purposes. 
To  be  prepared  for  such  an 
emergency,  many  engineers 
design  the  furnaces  for  coal 
burning  and  arrange  them 
with  loose  brick- work  for  the 
temporary  use  of  oil. 

Fig.  20  illustrates  the 
application  of  a  Hammel 
burner  at  the  rear  end  of  a 
furnace. 

Oil-Fired  Furnaces :  Engr.  U.S., 
July  1,  1902,  p.  491,  Nov.  15, 1905; 
Eng.  Mag.,  May,  1903;  Engng., 
April  15,  1904,  p.  523,  April  29, 
1904,  p.  594.  Liquid  Fuel  Com- 
bustion: Trans.  A.S.M.E.,  May, 
1902.  Equipment  for  Oil  Fuel: 
Eng.  and  Min.  Jour.,  Oct.  7, 
1905. 

53.  Air  vs.  Steam  as  an 
Atomizing  Medium.  —  Table 
10  gives  the  results  of  a 
series  of  tests  made  by  the 
U.S.  Naval  Liquid  Fuel  Board 
in  1902  on  different  types  of 
burners  using  air,  steam,  or 
both  for  atomizing  the  fuel. 
The  first  eight  tests  were 
made  with  compressed  air 

as  the  spraying  medium  and  under  pressures  varying  from  0.78  pounds 
to  4.68  pounds  per  square  inch.     The  most  economical   results  were 


FUELS  AND  COMBUSTION 


61 


obtained  with  the  lower  pressures,  and  the  total  steam  used  to  compress 
the  air  varied  from  1.06  per  cent  to  7.45  per  cent  of  the  total  steam 
generated,  but  not  all  burners  would  work  with  air  at  this  low  pres- 
sure. With'  steam  as  the  spraying  medium  the  steam  required  to 


FIG.  20.    Furnace  for  burning  Fuel  Oil,  Rear  Feed. 

operate  the  burner  varied  from  3.98  per  cent  to  5.77  per  cent  of 
the  total  generated,  while  the  burner  using  both  steam  and  air 
required  6.09  per  cent  to  8.54  per  cent  of  the  total.  The  results  of 
recent  tests  with  the  latest  types  of  burners  give  somewhat  lower 
steam  consumption  than  the  tests  conducted  by  the  Naval  Board 
and  a  good  average  is  not  far  from  3  per  cent.*  Table  10  contains 
also  the  results  of  a  few  scattering  tests  made  with  different  types  of 
burners. 

In  general  it  may  be  said  that  where  a  supply  of  low-pressure  air 
is  available,  air  is  unquestionably  more  economical  than  steam  as  an 
atomizing  medium,  but  in  the  average  boiler  plant  the  use  of  steam 
obviates  complication  and  risk  of  interrupted  service.  Where  it  is 
necessary  to  use  high-pressure  air  the  economy  decreases  with  the 
increase  in  pressure,  since  the  cost  of  each  cubic  foot  of  compressed  air 
increases  rapidly  with  the  pressure,  but  its  ability  to  atomize  the  oil 
does  not  increase  proportionately. 

Steam  vs.  Air  for  Liquid  Fuel  and  Oil  Burners  :  Am.  Mach.,  Vol.  27,  No.  51. 
*  See  Proceedings,  A.  S.  M.  E.,  Dec.,  1908,  p.  1698. 


62 


STEAM  POWER  PLANT  ENGINEERING 


PQ 


jad)  J9iiog 


•ftirao 


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spunoj)  ajnssaij 


oo  eo  »o  1-4  TJ«  os  oo  TP  os  t—  t>.  co  »-i  t*. 


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FUELS  AND  COMBUSTION 


63 


54.  Oil  Pressure.  —  This  varies  with  the  different  types  of  burners 
and  ranges  from  a  few  pounds  to  60  pounds  or  more  per  square  inch. 
The  low-pressure  systems  are  ordinarily  operated  under  standpipe  pres- 
sures as  in  Fig.  21,  which  illustrates  the  arrangement  of  apparatus  as 
advocated  by  the  International  Gas  and  Fuel  Company.  A  steam  pump 
B  draws  the  oil  from  the  buried  tank  through  pipe  Z  and  delivers  it  to 
the  standpipe  E.  Thence  it  flows  through  pipe  /  to  the  burners  under  a 


FIG.  21.    International  Gas  and  Fuel  Company's  Fuel  Oil  System. 

head  of  about  10  feet.  The  pump  runs  constantly,  the  surplus  oil  flow- 
ing back  to  the  tank  through  the  pipe  T.  The  oil  is  heated  by  the 
exhaust  pipe  Z'.  The  oil  pump  is  provided  with  a  device  D  having  a 
piston  connected  by  a  chain  with  a  cock  S,  which  automatically  opens 
when  the  boiler  is  not  under  steam  pressure,  so  that  the  standpipe  will 
be  emptied,  the  oil  flowing  to  the  storage  tank. 

The  high-pressure  systems  are  invariably  operated  by  steam  pumps, 
usually  in  duplicate,  and  are  so  arranged  that  the  oil  pressure  will  be 
kept  practically  constant  irrespective  of  the  steam  pressure.  The  adjust- 
ment of  steam  and  oil  is  a  very  delicate  operation,  and  fluctuation  in  the 
steam  pressure  disturbs  the  proportion  of  oil  and  steam ;  to  prevent  this 


64 


STEAM  POWER  PLANT  ENGINEERING 


the  steam  pressure  at  the  burner  is  reduced  several  pounds  below  that 
of  the  boiler  by  suitable  reducing  valves  and  is  thereby  kept  at  a  nearly 
constant  value. 

55.  Oil  Storage  and  Transportation.  —  Distillates  or  reduced  oils  are 
readily  stored  and  transported,  but  the  crude  oils,  on  account  of  the 
inflammability  of  the  highly  volatile  elements,  offer  a  different  problem. 
In  most  cities  distillates  may  be  stored  in  large  quantities  but  only  in 
tanks  sunk  below  the  lowest  level  of  the  surrounding  territory.  This 
is  a  protection  against  flooding  the  district  with  burning  oil  in  case  of 
a  fire.  In  the  country  the  oil  is  ordinarily  stored  in  tanks  above  the 
ground  level  and  at  some  distance  from  the  plant. 

Fig.  22  illustrates  the  Hydraulic  Oil  Storage  Company's  system  of 
storing  oil  and  delivering  it  to  the  burners.  The  oil  reservoirs  are  placed 


SIPHON   BREAKER 


OIL    TO    BURNERS 
DISCHARGE   TO   SEWER 


FIG.  22.    Hydraulic  Oil  Storage  Company's  Fuel  Oil  System." 


below  grade  as  indicated  to  minimize  fire  risk.  The  operation  is  as 
follows:  Water  enters  the  "  float  box  "  and  flows  through  a  "  three-way 
cock  "  to  the  bottom  of  the  reservoir  until  all  of  the  oil  and  water 
pipes  are  filled  up  to  the  level  of  the  float  box,  when  the  float  auto- 
matically cuts  off  the  supply.  This  flooding  of  the  entire  system 


FUELS  AND  COMBUSTION  65 

drives  out  all  of  the  air.  The  three-way  cock  is  then  turned  to 
"  discharge "  and  part  of  the  water  flows  to  the  sewer.  The  tank- 
car  or  wagon  is  next  attached  to  the  "  oil  inlet "  and  the  oil  flows  into 
the  tank  and  displaces  the  water  until  the  level  of  the  "  filler  float  " 
is  reached,  when  the  supply  is  automatically  cut  off.  The  inlet  is  so 
placed  that  the  head  of  oil  in  the  tank-car  is  sufficiently  great  to  over- 
come the  opposing  head  of  water.  The  three-way  valve  is  next  turned 
to  the  first  position  and  the  head  of  water  forces  the  oil  to  the  burners. 
After  the  oil  has  been  withdrawn  from  the  storage  tank  the  water  can 
only  rise  to  the  level  of  the  water  in  the  float  box  and  therefore  cannot 
be  fed  to  the  furnace.  The  small  steam  pipe  admits  steam  into  the 
tank  and  heats  the  oil,  thereby  making  it  flow  more  freely. 

Staring  Oil  Fuels :  Eng.  News,  Oct.  29,  1903,  p.  396. 

Petroleum  Reservoirs :  Jour.  Soc.  Chem.  Ind.,  Jan.  31,  1899. 

Handling  Fuel  at  Railway  Terminals  :  Eng.  News,  Sept.  25,  1902,  p.  232. 

56.  Conclusions  of  TJ.  S.  Naval  Liquid  Fuel  Board.  —  After  a  series 
of  elaborate  tests  it  was  concluded 

a.   That  oil  can  be  burned  in  a  very  uniform  manner. 

6.  That  the  evaporative  efficiency  of  nearly  every  kind  of  oil  per 
pound  of  combustible  is  probably  the  same.  While  the  crude  oil  may 
be  rich  in  hydrocarbons,  it  also  contains  sulphur,  so  that,  after  refining, 
the  distilled  oil  has  probably  the  same  calorific  value  as  the  crude 
product. 

c.  That  a  marine  steam  generator  can  be  forced  to  even  as  high  a 
degree  with  oil  as  with  coal. 

d.  That  up  to  the  present  time  no  ill  effects  have  been  shown  upon 
the  boiler. 

e.  That  the  firemen  are  disposed  to  favor  oil,   and   therefore   no 
impediment  will  be  met  in  this  respect. 

/.  That  the  air  requisite  for  combustion  should  be  heated  if  possible 
before  entering  the  furnace.  Such  action  undoubtedly  assists  the 
gasification  of  the  oil  product. 

g.  That  the  oil  should  be  heated,  so  that  it  could  be  atomized  more 
readily. 

h.  That  when  using  steam  higher  pressures  are  undoubtedly  more 
advantageous  than  lower  pressures  for  atomizing  the  oil. 

i.  That  under  heavy  forced  draft  conditions,  and  particularly  when 
steam  is  used,  the  Board  has  not  yet  found  it  possible  to  prevent 
smoke  from  issuing  from  the  stack,  although  all  connected  with  the  tests 


66  STEAM  POWER  PLANT  ENGINEERING 

made  special  efforts  to  secure  complete  combustion.  Particularly  for 
naval  purposes,  it  is  desirable  that  the  smoke  nuisance  be  eradicated  in 
order  that  the  presence  of  a  war  ship  might  not  be  detected  from  this 
cause.  As  there  has  been  a  tendency  of  late  to  force  the  boilers  of 
industrial  plants,  the  inability  to  prevent  the  smoke  nuisance  under 
forced-draft  conditions  may  have  an  important  influence  upon  the 
increased  use  of  liquid  fuel. 

j.  That  the  consumption  of  liquid  fuel  cannot  probably  be  forced 
to  as  great  an  extent  with  steam  as  the  atomizing  agent  as  when  com- 
pressed air  is  used  for  this  purpose.  This  is  probably  due  to  the  fact 
that  the  air  used  for  atomizing  purposes,  after  entering  the  furnace, 
supplies  oxygen  for  the  combustible,  while  in  the  case  of  steam 
the  rarefied  vapor  simply  displaces  air  that  is  needed  to  complete 
combustion. 

k.  That  the  efficiency  of  oil-fuel  plants  will  be  greatly  dependent 
upon  the  general  character  of  the  installation  of  auxiliaries  and  fittings, 
and  therefore  the  work  should  be  intrusted  only  to  those  who  have 
given  careful  study  to  the  matter  and  who  have  had  extended  experience 
in  burning  the  crude  product.  The  form  of  the  furnace  will  play  a  very 
small  part  in  increasing  the  use  of  crude  petroleum.  The  method  and 
character  of  the  installation  will  count  for  much,  but  where  burners  are 
simple  in  design  and  are  constructed  in  accordance  with  scientific 
principles  there  will  be  very  little  difference  in  their  efficiency.  Con- 
sumers should  principally  see  that  they  do  not  purchase  appliances  that 
have  been  untried  and  have  been  designed  by  persons  who  have  had 
but  limited  experience  in  operating  oil  devices. 

57.  Gaseous  Fuels.  —  These  fuels  offer  all  of  the  advantages  of  liquid 
fuels  and  but  few  of  the  disadvantages.  The  gases  most  commonly  met 
with  in  connection  with  steam  power  plants  are  outlined  in  Table  11. 
The  artificial  gases  for  steam  purposes  are  prohibitive  in  cost  in 
most  cases,  and  even  in  blast-furnace  installations,  where  the  gases  are 
waste  products,  the  gas  engine  has  virtually  supplanted  the  steam  engine 
for  power  purposes.  In  the  immediate  locality  of  natural  gas  wells  gas- 
fired  furnaces  may  prove  to  be  more  economical  than  coal  furnaces,  but 
the  limited  supply  and  constant  fluctuation  in  price  limit  its  use  as  a 
general  fuel.  From  the  market  quotations  on  coal  and  gas  and  the 
comparative  heating  value  of  each  the  actual  cost  per  B.T.U.  is  readily 
obtained,  and  by  combining  this  with  the  relative  efficiencies  from  the 
furnace  standpoint  the  net  cost  of  the  fuel  is  obtained.  The  following 
table,  based  upon  the  assumption  that  one  cubic  foot  of  natural  gas 
under  standard  conditions  has  a  heating  value  of  1,000  B.T.U. ,  will 
enable  an  approximate  comparison  to  be  made: 


FUELS  AND  COMBUSTION 


B.T.U.  per  Pound  of 
Coal 

Pounds  of  Coal  Equal 
to  1,000  Cu.  Ft. 

*No.  of  1,000  Cu.  Ft. 
of  Gas  Equal  to  One 

of  Gas. 

Short  Ton  of  Coal. 

10,000 

100 

20 

11,000 

91 

22 

12,000 

83 

24 

13,000 

77 

26 

14,000 

71 

28 

15,000 

67 

30 

Fuel  Economy  :  Fuel  Economy  in  Steam  Power  Plants :  Cassier's  Mag.,  May,  1904; 
Inst.  of  Elec.  Engrs.,  Jan.  12,  1905;  Engr.  U.S.,  April  1,  1905;  Engng.,  Aug.  7, 1903; 
Eng.  Mag.,  June,  1907.  The  Province  of  the  Fuel  Expert :  Eng.  and  Min.  Jour.,  May  25, 
1905.  A  Gas-Fired  Boiler :  Engr.  U.S.,  Feb.  15,  1907,  p.  223. 

See  also  A.S.M.E.  Code  for  conducting  Boiler  Tests  —  reprinted  in  Appendix  B. 

TABLE    11. 

CHARACTERISTICS    OF    GASEOUS    FUELS. 
(Lucke.) 


H 

CO 

CH4 

C2H4 

O 

Natural  gas  

1.7 
27.7 
39.78 
21.8 
49.50 
9.2 
14.0 
18.0 
21.0 
29.0 
53.0 
3.0 

0.55 
6.8 
7.04 
28.1 
35.93 
25.3 
20.0 
25.0 
12.0 
12.0 
6.0 
27.5 

94.16 
50.0 
45.16 
30.7 
1.05 
3.1 
2.0 
3.0 
2.0 
2.0 
35.0 

0.30 
13.0 
6.38 
12.9 

0.30 

"o.'oe" 

0.5 

Cannel-coal  gas  

Common-coal  gas  •  

Carburetted  water-gas  

Uncarburetted  water-gas  
Producer-gas,  little  steam  
Loomis  Pettibone  coal  gas  
Dowson  gas,  average  
Taylor  gas,  average  
Mond  gas     

0.8 
0.20 

0.10 
3.6 

Coke-oven  gas  

2.0 



Blast-furnace  gas  

C02 

N 

Cubic  Feet 
of  Air  per 
Cubic  Foot 
of  Gas. 

B.T.U.  per  Cubic 
Foot 
of  Gas. 

High. 

Low. 

Natural  gas 

0.29 
0.1 
1.08 
3.8 
4.25 
3.4 
8.2 
7.0 
6.0 
14.5 
2.0 
10.0 

2.80 
2.4 
0.50 
2.2 
8.75 
58.2 
55.5 
47.0 
57.0 
42.5 
2.0 
59.4 

9.13 
6.50 
6.38 
6.00 
2.10 
1.24 

989 
843 
727 
702 
295 
160 

888 
762 
651 
635 
265 
150 

Cannel-coal  gas 

Common-coal  gas 

Carburetted  water-gas 

Uncarburetted  water-gas  
Producer-gas,  little  steam  
Loomis  Pettibone  coal  gas  .... 
Dowson  gas,  average  
Taylor  gas,  average  

1.32 
0.98 
1.17 
5.06 
.81 

119 
130 
156 
620 
100 

115 
116 
139 
524 
99 

Mond  gas  

Coke-oven  gas  
Blast-furnace  gas  

CHAPTER  III. 

BOILERS. 

58.  As  affecting  fuel  economy  the  boiler  equipment  is  by  far  the 
most   important   part  of  the  power  plant  and  involves  the  largest 
share  of  the  operating   expenses.      It   matters  little  how  elaborate, 
modern,  or  well  designed  it  may  be,  skill,   good   judgment,  and  con- 
tinued vigilance  are  required  on  the  part  of  the  operator  to  secure  the 
best  efficiency. 

Of  the  various  types  and  grades  of  boilers  on  the  market  experience 
shows  that  most  of  them  are  capable  of  practically  the  same  evaporation 
per  pound  of  coal,  provided  they  are  designed  with  the  same  proportions 
of  heating  and  grate  surface  and  are  operated  under  similar  conditions. 
They  differ,  however,  with  respect  to  space  occupied,  weight,  capacity, 
first  cost,  and  adaptability  to  particular  conditions  of  operation  and 
location. 

59.  Classification.  —  As  to  design  and  construction  there  is  an  almost 
endless  variety  of  boilers   and  furnaces,   classified   as  internally  and 
externally  fired;  water  tube  and  fire  tube;  through  tube  and  return  tubular; 
horizontal  and  vertical. 

The  internally  fired  type  includes  the  vertical  tubular,  locomotive, 
Scotch-marine,  and  practically  all  flue  boilers.  The  externally  fired 
includes  the  plain  cylinder,  the  through  tubular,  return  tubular,  and 
nearly  all  stationary  water-tube  boilers. 

60.  Vertical  Tubular  Boilers.  —  Vertical  tubular  boilers,  Figs.  1  and 
23,  are  commonly  used  where  small  power,  compactness,  low  first  cost, 
and  sometimes  portability  are  the  chief  requirements,  though  they  are 
not  necessarily  restricted  to  small  sizes.     The  tubes  are  sometimes 
arranged  so  that  the  spaces  between  them  radiate  from  a  hand  hole  on 
one  side  so  that  a  scraper  may  readily  be  inserted  to  clean  the  top  of  the 
furnace  plate.     The  hand  hole  in  the  water  leg  permits  removal  of  the 
scale.     It  is  convenient  to  place  a  chain  in  the  bottom  of  the  water  leg 
which  can  be  worked  around  through  the  hand  hole  for  the  purpose 
of  loosening  up  the  scale  deposit.      The  distance  between  the  furnace 
crown  and  top  of  the  grate  is  never  less  than  24  inches  even  in  the  smallest 
boiler  and  should  be  as  great  as  possible  to  insure  good  combustion. 

/Two  styles  of  vertical  boilers  are  in  common  use,  the  ordinary  vertical 

68 


BOILERS 


69 


type,  Fig.  1,  and  the  submerged  type,  Fig.  23.  In  the  former  the  upper 
tube  sheet  and  part  of  the  tubes  are  above  the  water  line,  and  while  this 
feature  may  tend  to  superheat  the  steam  to  a  slight  extent,  the  difficulty 


STACK 


STEAM  GAUGE 


HAND-HOLE 


WATER  COLUMN 


STAY-BOLTS 


BLOW  OFF 


FIG.  23.    Vertical  Tubular  Boiler  with  Submerged  Tube  Sheet. 

from  unequal  expansion  and  liability  to  overheating  is  of  sufficient 
moment  to  justify  the  use  of  the  submerged  type,  particularly  where  the 
boiler  is  likely  to  be  forced  above  its  rated  capacity.  The  advantages 
of  this  type  of  boiler  are  (1)  compactness  and  portability;  (2)  requires 
no  setting  beyond  a  light  foundation;  (3)  is  a  rapid  steamer,  and  (4)  is 
low  in  first  cost.  The  disadvantages  are  (1)  inaccessibility  for  thorough 
inspection  and  cleaning;  (2)  small  steam  space,  which  results  in  excessive 


70 


STEAM  POWER  PLANT  ENGINEERING 


priming  at  heavy  loads;  (3)  poor  economy  except  at  light  loads,  as  the 
products  of  combustion  escape  at  a  high  temperature  on  account  of  the 


SECTION    THROUGH        A-B 


SECTION    THROUGH     ASH     PIT 


SECTIONAL  FRONT  ELEVATION 


FIG.  24.    Manning  Vertical 
Fire-Tube  Boiler. 


shortness  of  the  tubes;  (4)  smokeless  combustion  practically  impossible 
with  bituminous  coals;  (5)  the  small  water  capacity  results  in  rapidly 
fluctuating  steam  pressures  with  varying  demands  for  steam. 


BOILERS 


71 


Although  vertical  fire-tube  boilers  are  usually  of  very  small  size, 
being  seldom  constructed  in  sizes  over  60  horse  power,  an  exception  is 
found  in  the  Manning  boiler,  Fig.  24,  which  is  constructed  in  sizes  as 
large  as  250  horse  power.  Many  of  the  disadvantages  found  in  the 
smaller  types  are  obviated  in  the  Manning  boilers,  which,  as  far  as 
safety  and  efficiency  are  concerned,  rank  with  any  of  the  other  first- 
class  types.  They  differ  from  the  boiler  described  above  mainly  in 
having  the  lower  or  furnace  portion  of  much  greater  diameter  than  the 
upper  part  which  encircles  the  tubes.  This  permits  a  proper  proportion 
of  grate,  which  is  not  obtainable  in  boilers  like  Figs.  1  and  23.  The 
double  flanged  head  connecting  the  upper  and  lower  shells  allows 
sufficient  flexibility  between  the  top  and  bottom  tube  sheets  to  provide 
for  unequal  expansion  of  tubes  and  shell.  The  ash  pit  is  built  of  brick 
and  the  water  leg  does  not  extend  below  the  grate  level,  thus  doing  away 
with  dead  water  space.  Where  overhead  room  permits  and  ground 
space  is  expensive,  this  boiler  offers  the  advantage  of  taking  up  a  small 
floor  space  as  compared  with  horizontal  types. 

61.  Fire-Box  Boilers.  —  Although  vertical  fire-tube  boilers  may  be 
classed  as  fire-box  boilers,  yet  the  term  "  fire  box  "  is  usually  associated 
with  the  locomotive  types,  whether  used  for  traction  or  stationary  pur- 
poses. The  usual  form  of  fire-box  boiler  as  applied  to  stationary  work 


SAFETY  VALVE 


STAY 


STEAM 
DOME 


FIRE-TUBES 


STAY-BOLTS 


CROWN    SHEET 


GRATE 


FIRE   DOOR 


ASH   DOOR 


SADDLE     SUPPORT 


FIG.  25.    Typical  Fire-box  Boiler.  —  Stationary  Type. 

is  illustrated  in  Fig.  25.  The  shell  is  prolonged  beyond  the  front  tube 
sheet  to  form  a  smoke  box.  The  front  ends  of  the  tubes  lead  into  the 
smoke  box  and  the  rear  ends  into  the  furnace  or  fire  box.  The  fire  box 


72  STEAM  POWER  PLANT  ENGINEERING 

is  ordinarily  of  rectangular  cross  section,  and  is  secured  against  collapse 
by  stay  bolts  and  other  forms  of  stays.  In  Fig.  25  the  smoke  box  is  of 
cylindrical  cross  section  and  hence  requires  no  staying  except  at  the 
flat  surface.  Fire-box  boilers  are  used  a  great  deal  in  small  heating 
plants  where  space  limitation  precludes  other  types.  Their  steam 
capacity  gives  them  an  advantage  over  the  vertical  tubular  form. 
Being  internally  fired  no  brick  setting  is  required.  They  are  usually 
of  cheap  construction,  designed  for  low  pressure,  and  seldom  made  in 
sizes  over  75  horse  power.  Unless  carefully  designed  and  constructed 
high  steam  pressures  are  apt  to  cause  leakage  because  of  unequal  expan- 
sion of  boiler  shell,  tubes,  and  fire  box.  Portable  fire-box  boilers  with 
return  tubes  are  made  in  sizes  as  large  as  150  horse  power  and  for 
pressures  as  high  as  150  pounds  per  square  inch,  but  being  more  costly 
than  some  of  the  other  types  of  boilers  of  equal  capacity  are  used  only 
where  portability  is  an  essential  requirement. 

62.  Scotch-Marine  Boiler.  —  Where  an  internally  fired  boiler  is 
desired  for  large  powers  the  Scotch-marine  type  is  finding  much  favor 
with  engineers.  A  number  of  the  tall  office  buildings  in  Chicago  are 
equipped  with  boilers  of  this  class  which  are  giving  good  results.  They 
require  little  overhead  room,  no  brick  setting,  and  are  excellent  steamers. 


FIG.  26.    Stationary  Scotch-Marine  Boiler. 

The  Continental  boiler,  Fig.  26,  is  one  of  the  best  known  of  this  type. 
The  boiler  is  self-contained  and  requires  no  brick  setting,  the  only  fire 
brick  used  being  those  that  form  the  bridge  wall,  baffle  ring  and  the 
layer  at  the  back  of  the  combustion  chamber.  The  furnace  and  tubes  are 


BOILERS 


73 


entirely  surrounded  by  water,  so  that  all  fire  surfaces,  excepting  the 
rear  of  the  combustion  chamber,  are  water  cooled.  The  furnace  is  cor- 
rugated for  its  whole  length.  These  corrugations,  in  addition  to  giving 
greater  strength  to  the  furnace,  act  as  a  series  of  expansion  joints, 
taking  up  the  strains  due  to  unequal  expansion  of  furnace  and  shell. 
Practically  all  types  of  mechanical  stokers  and  grates  are  applicable  to 
these  boilers.  The  advantages  of  a  Scotch  boiler  and  of  all  internally 
fired  boilers  are  (1)  minimum  radiation  losses;  (2)  requires  no  setting; 
(3)  no  leakage  of  cool  air  into  the  furnace  as  sometimes  occurs  through 
cracks  or  porous  brickwork  of  other  types;  (4)  large  steaming  capacity 
for  the  space  occupied.  The  circulation,  however,  is  not  always  positive 
and  the  water  below  the  furnace  may  be  considerably  below  the  average 
or  normal  temperature,  giving  rise  to  unequal  expansion  and  contraction 
which  may  cause  leakage.  The  boiler  proper  is  relatively  costly,  but 
this  is  offset  to  some  extent  by  the  absence  of  setting. 

63.   Robb-Mumford   Boiler.  —  Fig.    27   shows   a   section   through   a 
Robb-Mumford  boiler,  which  is  a  modification  of  the  Scotch-marine 


FIG.  27.    Robb-Mumford  Boiler. 

and  of  the  horizontal  tubular  type.  It  consists  of  two  cylindrical 
shells,  the  lower  one  containing  a  round  furnace  and  tubes  and  the 
upper  one  forming  the  steam  drum,  the  two  being  connected  by  two 
necks.  The  lower  shell  has  an  incline  of  about  one  inch  per  foot  from 
the  horizontal,  for  the  purpose  of  promoting  circulation  and  draft, 
and  also  for  convenience  in  washing  out  the  lower  shell.  Combustion 
takes  place  in  the  furnace,  which  is  surrounded  entirely  by  water,  and 


74  STEAM  POWER  PLANT  ENGINEERING 

the  gases  pass  through  the  tubes  and  return  between  the  lower  and 
upper  shells  (this  space  being  enclosed  by  a  steel  casing)  to  the  outlet 
at  the  front  of  the  boiler.  Mingled  water  and  steam  circulate  rapidly 
up  the  rear  neck  into  the  steam  drum,  where  the  steam  is  released,  the 
water  passing  along  the  upper  drum  towards  the  front  of  the  boiler  and 
down  the  front  neck,  a  semicircular  baffle  plate  around  the  furnace 
causing  the  down-flowing  water  to  circulate  to  the  lowest  part  of  the 
lower  shell  under  the  furnace.  The  outer  casing,  which  incloses  the  space 
between  the  lower  and  upper  shells,  including  the  rear  smoke  box  and 
the  smoke  outlet,  is  constructed  of  steel  plate,  with  angle-iron  stiffeners, 
the  various  sections  being  bolted  together  for  convenient  removal.  The 
inside  of  the  steel  case,  including  the  rear  smoke  chamber,  is  lined  with 
asbestos  air-cell  blocks  fitted  in  between  the  angle-iron  stiffeners.  The 
top  of  the  upper  drum  and  bottom  of  the  lower  shell  are  also  covered 
with  non-conducting  material  after  the  boiler  is  erected.  Owing  to 
the  fact  that  steam  and  water  spaces  are  divided  between  two  cylin- 
drical shells,  the  thickness  of  plates  is  not  so  great  as  in  the  Scotch- 
marine  or  horizontal  return  tubular  types;  and  the  rear  chamber  of  the 
marine  boiler  is  avoided. 

The  chief  claim  for  this  type  of  boiler  is  compactness.  A  battery  of 
five  200-horse-power  units  occupies  a  floor  space  of  but  33  feet  in  width 
by  20  feet  in  depth  and  12.5  feet  high.  Each  unit  is  entirely  independent 
and  may  be  isolated  for  cleaning,  inspection,  and  repairs. 

64.  Horizontal  Return  Tubular  Boilers.  —  These  are  the  most  common 
in  use  and  are  constructed  in  sizes  up  to  200  horse  power.  They  are 
simple  and  inexpensive  and,  when  properly  operated,  durable  and 
economical.  Figs.  28  to  31  show  various  forms  of  standard  settings, 
and  Figs.  75,  76,  and  76a  different  "  smokeless  "  settings.  The  grate  is 
independent  of  the  boiler,  and  the  products  of  combustion  pass  beneath 
the  shell  to  the  back  end,  returning  through  the  tubes  to  the  front, 
and  into  the  smoke  connection. 

The  tubes  are  3  to  4  inches  in  diameter  and  from  14  to  18  feet  long, 
and  are  expanded  into  the  tube  sheets.  The  portion  of  the  tube  sheets 
not  supported  by  the  tubes  is  secured  against  bulging  by  suitable  stays. 
Access  to  the  interior  of  the  boiler  is  obtained  through  manholes.  The 
most  convenient  arrangement  for  inspection  and  cleaning  is  to  have 
one  manhole  located  at  the  top  of  the  shell  and  one  at  the  bottom  of 
the  front  tube  sheet.  Return  tubular  boilers  are  made  either  with  an 
extended  front  (Fig.  28)  or  flush  front  (Fig.  29).  The  latter  costs  a 
little  more  for  brick  and  setting,  but  it  is  more  convenient  to  operate 
and  the  boiler  is  less  expensive.  The  shell  may  be  supported  by  lugs  on 
the  brickwork  as  in  Fig.  28  or  by  steel  beams  and  hangers  as  in  Fig,  30. 


BOILERS 


75 


76 


STEAM  POWER  PLANT  ENGINEERING 


BOILERS 


77 


The  latter  construction  permits  the  brickwork  and  shell  to  expand 
or  contract  independently,  and  settling  of  the  brickwork  does  not  affect 
the  boiler  alignment.  With  the  side  bracket  support,  the  front  lugs 
usually  rest  directly  on  iron  or  steel  plates  imbedded  in  the  brickwork, 
and  the  back  lugs  on  rollers,  to  permit  free  expansion  and  contraction. 
The  brackets  are  long  enough  to  rest  upon  the  outside  wall,  so  that  the 
inside  brick  lining  can  be  renewed  without  disturbing  the  setting.  The 
distance  between  the  rear  tube  sheet  and  wall  should  be  about  16  inches 


FIG.  30.    Return  Tubular  Boiler  Setting.  —  Steel  Beam  Suspension. 

for  boilers  less  than  60  inches  in  diameter  and  from  20  to  24  inches  for 
larger  ones.  The  distance  between  grate  and  boiler  shell  should  not 
be  less  than  28  inches  for  anthracite  coal  and  36  inches  for  bituminous 
coal.*  The  greater  this  distance  the  more  complete  the  combustion, 
since  the  gases  will  have  a  better  opportunity  for  combining  with  the  air 
before  coming  in  contact  with  the  comparatively  cool  surfaces  of  the 
shell.  The  shell  should  be  slightly  inclined  toward  the  blow-off  end 
so  as  to  drain  freely. 

The  vertical  distance  between  the  bridge  wall  and  shell  is  usually 
between  10  and  12  inches.  The  lower  part  of  the  combustion  chamber 
behind  the  bridge  wall  may  be  filled  with  earth  and  paved  with  common 

*  For  smokeless  combustion  the  setting  must  be  modified.  See  furnace  illustrated  and 
described  in  paragraph  96. 


78 


STEAM  POWER  PLANT  ENGINEERING 


•c 


I 


BOILERS 


79 


brick,  as  in  Fig.  31,  or  left  empty  as  in  Fig.  29.  The  shape  of  the  walls, 
whether  curved  to  conform  to  the  shell  or  flat,  appears  to  have  little 
influence  on  the  economy. 

The  side  and  end  walls  are  ordinarily  constructed  of  common  brick 
with  an  inner  lining  of  fire  brick,  and  may  be  solid  as  in  Fig.  29  or 
double  with  air  spaces  as  in  Fig.  28.  The  latter  construction  is  prefer- 
able and  permits  the  inner  and  outer  walls  to  expand  independently 
without  cracking  and  settling.  The  side  walls  are  braced  by  five  pairs 
of  buck-staves,  with  through  rods  under  the  paving  and  over  the  tops 
of  the  boilers. 

The  connection  between  the  rear  wall  and  the  shell  is  a  source  of 
more  or  less  trouble  on  account  of  the  expansion  and  contraction  of  the 
boiler.  Cast-iron  supports  of  T  section  supporting  a  fire-brick  arch 
are  usually  employed  as  illustrated  in  Fig.  32,  the  clearance  between 
the  arch  and  the  shell  being  sufficient  to  allow  the  necessary  expansion. 


FIG.  32.    Furnace  Arch  Bars. 


FIG.  33.    Back  connection  made  with 
Cast-iron  Plate. 


Fig.  33  shows  the  common  method  of  resting  one  end  of  the  arch 
supports  on  the  rear  wall  and  the  other  end  on  an  angle  iron  riveted  to 
the  boiler. 

The  products  of  combustion  are  sometimes  carried  over  the  top  of  the 
boiler  as  shown  in  Fig.  31.  This  tends  to  superheat  the  steam,  but  the 
advantage  gained  is  probably  offset  considerably  by  the  extra  cost  of 
the  setting  and  the  accumulation  of  soot  on  the  top  of  the  shell.  The 
arrangement  is  not  common. 

The  steam  connection  is  naturally  made  to  the  highest  point  in  the 
boiler  shell.  Frequently  a  steam  dome,  to  which  the  steam  nozzle  is 
connected,  is  provided  as  in  Fig.  29.  The  function  of  the  steam  dome  is 
to  increase  the  steam  space  so  as  to  permit  the  collection  of  dry  steam  at 
a  point  high  above  the  water  level.  If  a  boiler  is  too  small  for  its  work 


80 


STEAM  POWER  PLANT  ENGINEERING 


BOILERS 


81 


and  is  forced  far  above  its  rating  a  steam  dome  is  probably  an  advantage, 
though  its  use  is  less  common  now  than  formerly,  since  a  properly 
designed  boiler  insures  ample  steam  space  without  one.  A  dry  pipe 
inside  the  boiler  above  the  water  line  as  in  Fig.  26  or  27  is  commonly 
used  to  guard  against  priming  where  the  nozzle  is  connected  to  the  shell. 

For  low  pressures  and  small  powers  the  return  tubular  boiler  has  the 
advantage  of  affording  a  large  heating  surface  in  a  small  space  and 
large  overload  capacity.  It  requires  little  overhead  room  and  its  first 
cost  is  low.  On  the  other  hand  the  interior  is  difficult  of  access  for 
purposes  of  cleaning  and  inspection.  Boilers  of  this  type  are  seldom 
constructed  in  sizes  above  150  horse  power  or  for  pressures  over  150 
pounds  per  square  inch,  since  the  cost  increases  rapidly  as  the  pressure 
rises  above  this  amount. 

65.  Babcock  &  Wilcox  Boiler.  —  Fig.  34  shows  a  longitudinal  section 
through  a  Babcock  &  Wilcox  boiler,  illustrating  a  typical  horizontal 
water-tube  type.  The  tubes,  usually  4  inches  in  diameter  and  18  feet 


FIG.  35.  Details  of 
Header, — Babcock 
and  Wilcox  Boiler. 


FIG.  36.    Front  Section,  —  Babcock  and  Wilcox  Boiler. 

in  length,  are  arranged  in  vertical  and  horizontal  rows  and  are  expanded 
into  pressed-steel  headers.  Two  vertical  rows  are  fitted  to  each  header 
and  are  "  staggered  "  as  shown  in  Fig.  35.  The  headers  are  connected 


82  STEAM  POWER  PLANT  ENGINEERING 

with  the  steam  drum  by  short  tubes  expanded  into  bored  holes.  Each 
tube  is  accessible  for  cleaning  through  openings  closed  by  covers  with 
ground  joints  held  in  place  by  wrought-iron  clamps  and  bolts.  The 
tubes  are  inclined  at  an  angle  of  about  22  degrees  with  the  horizontal. 
The  rear  headers  are  connected  at  the  bottom  to  a  cast-iron  mud  drum. 
The  steam  drum  is  horizontal  and  the  headers  are  arranged  either  ver- 
tically or  at  right  angles  to  the  tubes.  The  boiler  is  supported  by  steel 
girders  resting  on  suitable  columns  independent  of  the  brick  setting. 
The  grate  is  placed  under  the  higher  ends  of  the  tubes,  the  products  of 
combustion  passing  at  right  angles  to  the  tubes  and  being  deflected 
back  and  forth  by  fire-tile  baffles.  The  feed  water  enters  the  front  of 
the  steam  drum  as  shown  in  Fig.  36.  A  rapid  circulation  is  effected  by 
the  difference  in  density  between  the  solid  column  of  water  in  the  rear 
header  and  the  mixed  steam  and  water  in  the  front  one.  B.  &  W. 
boilers  under  150  horse  power  have  but  one  steam  drum,  and  the  larger 
sizes  have  two.  The  number  of  tubes  varies  with  the  size  of  boiler, 
ranging  from  6  in  width  and  9  in  height  in  the  100-horse-power  boilers 
to  14  high  and  18  wide  in  the  500-horse-power  boilers. 

66.  Heine  Boiler.  —  Fig.  37  shows  a  longitudinal  section  through  a 
Heine  horizontal  water-tube  boiler.     This  boiler  differs  from  the  B.  &  W. 
boiler  in  that  the  tubes  are  expanded  into  a  single  large  header  con- 
structed of  boiler  steel.     The  drum  and  tubes  are  parallel  with  each 
other  and  inclined  about  22  degrees  with  the  horizontal.     The  feed 
water  enters  at  the  front  of  the  steam  drum  and  flows  into  the  mud 
drum,  from  which  it  passes  to  the  rear  header.     Steam  is  taken  from 
the  front  of  the  steam  drum  and  is  partially  freed  from  moisture  by  the 
dry  pipe  A.     A  baffle  over  the  front  header  prevents  an  excess  of  water 
from  being  carried  into  the  dry  pipe.     As  the  rear  header  forms  one  large 
chamber,  no  additional  mud  drum  is  necessary  and  the  sediment  is 
blown  off  from  the  bottom  by  the  blow-off  cock.     The  circulation  is 
somewhat  freer  than  in  the  B.  &  W.  boiler  on  account  of    the  large 
sectional  area  through  the  headers. 

67.  Wickes  Boiler.  —  Fig.   38   shows   a  section   through   a   Wickes 
vertical  boiler,  illustrating  the  vertical  water-tube  type.     The  steam 
drum  and  water  drum  are  arranged  one  directly  above  the  other.     The 
tubes  are  expanded  and  rolled  into  both  tube  sheets  and  are  divided 
into  two  sections  by  fire-brick  tile.     The  water  line  in  the  steam  drum 
is  carried  about  two  feet  above  the  tube  sheet,  leaving  a  space  of  five 
feet  between  water  line  and  top  of  the  drum.     This  affords  a  large 
steam   space   and  disengagement    surface.     Feed  water  is  introduced 
into  the  steam  drum  below  the  water  line  and  flows  downward  through 
the  tubes  of  the  second  compartment.     The  boiler  is  supported  by  four 


BOILERS 


83 


84 


STEAM  POWER  PLANT  ENGINEERING 


brackets  riveted  to  the  shell  of  the  bottom  drum  and  is  independent 
of  the  setting.  The  entire  boiler  is  enclosed  in  brickwork  and  is  com- 
pletely surrounded  by  the  products  of  combustion.  The  upper  part 


23 


El   B  i  0 


i   B   i 


FIG.  38.    Wickes  Vertical  Water  Tube  Boiler. 

of  the  steam  drum  acts  as  a  superheating  surface  and  tends  to  dry 
the  steam.  Wickes  boilers  are  simple  in  design,  easy  to  inspect  and 
clean,  low  in  first  cost,  and  comparable  in  efficiency  with  any  water- 
tube  type  of  boiler. 


BOILERS 


85 


86 


STEAM  POWER  PLANT  ENGINEERING 


67a.  Parker  Boiler.  —  Fig.  38a  shows  a  longitudinal  sectional  ele- 
vation and  an  end  sectional  elevation  of  a  1200-H.P.  Parker  Down- 
Flow  Boiler  with  double-ended  setting.  This  type  of  boiler  is  finding 
much  favor  with  engineers  for  central  stations  where  large  units  are 
desired.  The  Parker  boiler  differs  from  the  conventional  horizontal 
water-tube  boiler  principally  in  circulation  and  flexibility. 

Feed  water  is  pumped  into  the  economizer  or  feed  element  (1),  Fig. 
38a,  at  0,  0,  and  flows  downward  through  a  series  of  tubes,  discharging 
finally  into  the  drum  through  an  upcast  H.     In  a  large  unit,  as  illus- 
trated here,  there  are  two  feed  elements  and  two  drums.     The  circula- 
tion in  the  feed  element 
is    indicated    by    solid 
lines  and  arrow  points 
at   the  left   of  the  end 
sectional  elevation,  the 
tubes  having  been  omit- 
ted  from   the   drawing 
for  the  sake  of  clearness. 
The  intermediate  ele- 
ments   (2)    take    their 
water  supply  from  the 
bottom    of    the    drum 
through  a  cross-box  V, 
the     circulation    being 
downward,  as  indicated 
by      arrow      points, 

through  four  tube  wide  elements,  and  finally  discharge  it  through  an 
upcast  X  into  the  steam  space  of  the  drum.  Each  element  has  a 
"  down-comer  "  and  an  upcast.  In  the  smaller  sized  boilers  the  inter- 
mediate elements  are  omitted. 

The  evaporator  elements  (3)  take  their  water  supply  from  the  bottom 
of  the  drum  at  V,  the  circulation  being  downwards  through  two  tube 
wide  elements,  and  finally  discharge  it  into  the  drum  at  U.  The  last 
two  passes  of  the  water  are  through  the  two  bottom  tubes  of  each  ele- 
ment, thus  assuring  dry  steam  without  the  use  of  dry  pipes.  To  prevent 
reversal  of  flow  each  element  is  fitted  with  a  check  valve  at  the  admission 
end.  Each  drum  is  equipped  with  a  diaphragm,  as  indicated,  separating 
the  steam  and  water  spaces,  thus  insuring  against  foaming  and  priming. 
Saturated  steam  is  taken  from  the  drum  at  A  and  passes  by  way  of 
B  to  C,  where  it  enters  the  superheater  S.  The  superheated  steam 
leaves  the  superheater  at  D  and  passes  by  way  of  E  and  R  to  the  storage 
drum  N,  finally  leaving  the  boiler  at  G.  The  superheater  is  designed 


oiler  H.P. 

3  f- 

1 

1 

1 

BOILER  R 
&ZE?  OF 

OOM 
BOI 

AREAF 
LERS,2  1 

OR  VARIOUS  TYPES  ANE 
JOlliERS'lN  A  BATTERY 

\ 

1 

1  1  1 

\ 

B  4  W.3,5,T,8,9>HI,11>12>1S>U>15I16 
SterIing-2,4,6 
Heine-1 
Parker-20,11 
Hon»8l>y-18,l» 

\ 

Area  of  Battery  8q.  Ft.  per  E 

>  io  **.  b,  fe  '-*  o>  *« 

\ 

5 

5     , 

0 
4 

1     l 

> 

S 

s" 

15 

e  • 

\ 

12 
7 

13   " 

"> 

\ 

18 

•^  

•^^. 

—  —  ^ 

.1 

500                               1000                          „  1500                             20C 

Capacity  of  Battery  Boiler  H.P. 
FlG.  38b. 


BOILERS 


87 


to   maintain  an  approximately  constant  degree  of  superheat  for  all 
variations  in  load. 

All  tubes  are  connected  by  malleable-iron  junction  boxes,  the  interior 
of  each  tube  being  accessible  through  hand  holes  placed  opposite  the 


FIG.  39.     Stirling  Boiler  and  Setting. 

end  of  each  tube.  The  hand-hole  cover  plates  are  on  the  inside  of  the 
box  and  have  conical  ground  joints,  thus  dispensing  with  gaskets. 

The  Parker  boiler  is  built  single  or  double  ended,  with  or  without 
superheater,  and  in  sizes  ranging  from  50  H.P  to  2500  H.P.  standard 
rating. 

68.  Stirling  Boiler.  —  Fig.  39  shows  a  longitudinal  section  through 
a  Stirling  water-tube  boiler,  which  differs  considerably  from  the  types 
just  described.  Three  horizontal  steam  drums  and  one  horizontal  mud 
drum  are  connected  by  a  series  of  inclined  tubes.  The  tubes  are  bent 
at  the  ends  to  permit  them  to  enter  the  drums  radially.  Short  tubes 


88  STEAM  POWER  PLANT  ENGINEERING 

connect  the  steam  spaces  of  all  the  upper  drums  and  also  the  water 
spaces  of  the  front  and  middle  drums.  Suitably  disposed  fire-tile 
baffles  between  the  banks  of  tubes  direct  the  gases  in  their  proper 
course.  The  boiler  is  supported  on  a  structural  steel  framework  in- 
dependent of  the  setting.  The  feed  water  enters  the  rear  upper  drum, 
which  is  the  cooler  part  of  the  boiler,  and  flows  to  the  bottom  or  mud 
drum,  where  it  is  heated  to  such  an  extent  that  many  of  the  impurities 
are  precipitated.  There  is  a  rapid  circulation  up  the  front  bank  of 
tubes  to  the  front  drum,  across  to  the  middle  drum,  and  thence  down 
the  middle  bank  of  tubes  to  the  mud  drum.  The  interior  of  the  drums 
is  accessible  for  cleaning  by  manholes  located  in  the  ends.  The  Stirling 
furnace  is  distinctive  in  design.  A  fire-brick  arch  is  sprung  over  the 
grates  immediately  in  front  of  the  first  bank  of  tubes.  The  large  tri- 
angular space  between  boiler  front,  tubes,  and  mud  drum  forms  the 
combustion  chamber.  Stirling  boilers  are  somewhat  lower  in  first 
cost  than  other  types  of  water-tube  boilers  on  account  of  the  absence 
of  numerous  hand  holes  and  the  like  which  are  necessary  in  the  hori- 
zontal type. 

69.  Unit  of  Evaporation.  —  The  performance  of  a  boiler  and  furnace 
may  be  expressed  in  terms  of  the  weight  of  water  evaporated  per  hour 
per  square  foot  of  heating  surface  or  of  the  weight  evaporated  per  pound 
of  fuel.  To  reduce  all  performances  to  an  equal  basis  so  as  to  facilitate 
comparison  the  evaporation  under  actual  conditions  is  conveniently 
referred  to  the  equivalent  evaporation  from  a  feed-water  temperature 
of  212  degrees  F.  to  steam  at  atmospheric  pressure.  The  heat  required 
to  evaporate  one  pound  of  feed  water  at  a  temperature  of  212  degrees  F. 
into  steam  of  the  same  temperature,  or  "  from  and  at  212  degrees  " 
as  it  is  commonly  called,  is  965.7  B.T.U.*  The  ratio  of  the  heat  neces- 
sary to  evapprate  one  pound  of  water  under  actual  conditions  of  feed 
temperature  and  steam  pressure  to  the  heat  required  to  evaporate  one 
pound  from  and  at  212  degrees  is  called  the  factor  of  evaporation.  Thus 
for  dry  steam, 


-  ni. 

965.7 


in  which 


F  —  factor  of  evaporation. 

\  =  total  heat  of  one  pound  of  steam  at  observed  pressure. 
t  =  temperature  of  the  feed  water,  degrees  F. 

*  Recent  redeterminations  of  the  properties  of  saturated  steam  give  this  figure 
as  970.4. 


BOILERS 


89 


If  the  steam  is  wet, 


xr  +  q, 


(12) 


(13) 


'Dry  Burfcoes, 


WetfSurfM* 


in  which 

q  =  heat  in  liquid  at  observed  pressure. 

x  =  the  quality  of  the  steam. 

r  =  latent  heat  of  evaporation  at  observed  pressure. 

If  the  steam  is  superheated, 

A  =  r  +  q  +  Cts, 
in  which 

C  =  the  specific  heat  of  the  superheated  steam. 

ts  =  the  degree  of  superheat,  degrees  F. 

69a.  Heat  Transmission.  —  Fig.  39a  shows  a  section  through  a  boiler 
heating  plate  and  serves  to  illustrate  the  accepted  theory  of  heat  trans- 
mission. The  outer  surface  of 
the  plate  is  covered  with  a  thin 
layer  of  soot  and  a  film  of  gas, 
and  the  inner  surface  is  similarly 
protected  by  a  layer  of  scale  and 
a  film  of  steam  and  water.  It  is 
therefore  reasonable  to  assume 
that  the  dry  surface  of  the  plate 
is  located  somewhere  within  the 
film  of  gas,  and  the  wet  surface 
within  the  film  of  water  and 
steam. 

The  heat  is  imparted  to  the  dry 
surface  by  (1)  radiation  from  the 
hot  fuel  bed  and  furnace  walls, 
and  by  (2)  convection  from  the 
moving  furnace  gases.  The  heat 
is  transferred  through  the  boiler 
plate  and  its  coatings  purely  by 
conduction.  The  final  transfer 
from  the  wet  surface  to  the  boiler 
is  mainly  by  convection. 

Radiation  depends  on  the  temperature,  and  according  to  the  law  of 
Stephen  and  Boltzmann  is  approximately  proportional  to  the  difference 
between  the  fourth  power  of  the  absolute  temperature  of  the  fuel  bed 
and  furnace  walls  and  the  temperature  of  the  dry  surface  of  the  heating 
plate.  According  to  this  law  the  heat  transmitted  by  radiation  increases 
rapidly  with  the  increase  in  furnace  temperature.  In  the  modern  boiler 
the  surface  exposed  to  radiation  is  only  a  small  portion  of  the  total 


A  =  Average  Temperature  of. Moving  Gases. 
B  —  Average  Temperature  of  Dry  Surface. 
C  =  Average  Temperature  of  Wet  Surface. 
D  temperature  of  Water  in  Boiler. 


FIG.  39a. 


Heat  Transmission  through 
Boiler  Plate. 


90  STEAM  POWER  PLANT  ENGINEERING 

heating  surface,  and,  since  in  well-operated  furnaces  the  temperature  of 
the  furnace  cannot  be  increased  materially  on  account  of  practical  con- 
siderations, there  is  little  hope  of  increasing  the  capacity  of  a  boiler  by 
increasing  the  furnace  temperature. 

The  heat  imparted  to  a  boiler  plate  by  convection  may  be  determined 
by  the  following  equation  (Prof.  Perry,  "  The  Steam  Engine/'  1906 
Ed.,  p.  588): 

H=C(ti-  t2)  vd,  (13a) 

in  which 

H  =  B.T.U.  transferred  per  hour  per  sq.  ft.  of  heating  surface. 

C  =  a  coefficient  determined  by  experiment. 

Zi  =  temperature  of  the  moving  gases,  degrees  F. 

t2  =  temperature  of  the  dry  plate  surface,  degrees  F. 

v  =  velocity  of  the  gases,  feet  per  sec. 

d  =  density  of  the  gases,  Ibs.  per  cubic  foot. 

Prof.  Nicholson  gives  the  following  modifications  of  formula  (13a) 
as  applied  to  boiler  tubes  or  flues  (Engr.  Lond.,  Feb.  19,  1908): 


'-^d>  (13b> 

in  which 

t  =  mean  film  temperature. 

m  =  hydraulic  radius  =  area  of  tube  in  square  inches  -f-  perimeter 
of  the  tube  in  inches,  other  notations  as  in  (13a). 

Both  equations  are  based  upon  the  same  general  law  except  that  the 
latter  gives  a  means  of  determining  coefficient  C  in  terms  of  the  mean 
film  temperature  and  the  dimensions  of  the  flues  or  tubes. 

An  examination  of  equation  (13a)  shows  that  for  a  given  set  of  condi- 
tions the  heat  imparted  to  a  unit  of  dry  surface  of  heating  plate  varies 
directly  as  the  difference  between  the  temperature  of  the  hot  gases  and 
that  of  the  dry  surface  and  directly  as  the  velocity  and  density  of  the 
gases.  However,  the  density  of  the  gases  drops  with  the  rise  of  tem- 
perature, and  increase  in  furnace  temperature  does  not  necessarily  imply 
increase  in  heat  impartation.  It  is  the  utilization  of  the  velocity  factor, 
then,  which  offers  a  possibility  of  increasing  boiler  capacity  and  efficiency. 

Experiments  by  Prof.  Nicholson  and  the  U.  S.  Geological  Survey  show 
that  by  establishing  a  powerful  scrubbing  action  between  the  gases  and 
the  boiler  plate  the  protecting  film  of  gas  is  torn  off  as  rapidly  as  it  is 
formed  and  new  portions  of  the  hot  gases  are  brought  into  contact  with 
the  plate,  thereby  greatly  increasing  the  rate  of  heat  transmission. 
Similarly  the  faster  the  circulation  of  the  water  the  greater  will  be  the 


BOILERS 


91 


scrubbing  action  tending  to  remove  the  bubbles  of  steam  from  the  wet 
surface  and  the  more  rapid  will  be  the  transfer  from  the  plate  to  the 
boiler  water. 

The  resistance  of  the  metal  itself  is  so  small  that  it  may.be  neglected 
in  calculating  the  heat  trans- 
mission, and  it  may  be  logically 
assumed  that  the  plate  will  take 
care  of  all  the  heat  that  reaches 
its  dry  surface. 

Prof.  Nicholson  found  that  by 
filling  up  the  flue  of  a  Cornish 
boiler  with  an  internal  water 
vessel,  leaving  an  annular  space 
of  only  1  inch  around  the  latter, 
an  evaporation  eight  times  the 
ordinary  rate  was  effected  at  a 
flow  of  gases  330  feet  per  second 
(8  to  10  times  the  average  flow). 
The  fan  for  creating  the  draft 
consumed  about  4J%  of  the 
total  power. 

The  conclusion  is  that  the 
heating  surface  for  a  given  evap- 
oration at  the  present  rating 
may  be  reduced  as  much  as  90% 
for  the  same  output,  with  a  cor- 
responding reduction  in  the  size, 
cost  and  space  requirements,  or 
with  a  given  heating  surface  of 
standard  rating  the  output  may 
be  enormously  increased;  also 
the  increase  in  power  necessary 
to  create  the  draft  is  by  no 
means  comparable  with  the  ad- 
vantages gained. 

The  modern  locomotive  boiler 
is  the  nearest  approach  to  these 
conditions  in  practice.  Here  a  powerful  draft  forces  the  heated  gases 
through  small  tubes  at  a  very  high  velocity  and  an  enormous  evapo- 
ration is  effected  with  a  comparatively  small  heating  surface.  See 
Fig.  39b  for  influence  of  draft  on  the  capacity  of  a  torpedo  boat  boiler 
(Power  and  Engr.,  May  24,  1910). 


12 

z  .  .  /  . 

11  \\ 

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0  100  200 

Per  cent,  of  Rated  Capacity  Developed  by  Boiler. 

Power 

FIG.  39b.  Influence  of  Draft  on  the  Capacity 
of  a  Normand  Water-Tube  Boiler  on  the 
U.  S.  Torpedo  Boat  "  Diddle." 


92  STEAM  POWER  PLANT  'ENGINEERING 

These  principles  have  been  applied  to  a  limited  extent  to  stationary 
boilers  already  installed  by  making  the  gas  passages  smaller  as  compared 
to  the  length  by  means  of  suitable  baffles  (Fig.  38a)  and  by  forcing 
larger  weights  of  gas  through  the  boiler,  either  by  forced  draft  or  by 
increasing  the  grate  area  (Fig.  68a). 

In  a  general  sense  when  the  capacity  of  a  boiler  is  doubled  or  tripled 
the  over-all  efficiency  of  the  whole  steam-generating  apparatus  drops, 
but  the  advantage  gained  usually  offsets  the  loss  in  fuel  economy.  A 
close  examination  of  the  results,  however,  will  show  that  the  loss  in 
efficiency  is  due  more  to  low  furnace  efficiency  than  to  inability  of  the 
boiler  to  absorb  the  heat  generated. 

In  view  of  recent  experiments  it  is  not  unlikely  that  within  the  next 
ten  years  boilers  will  be  constructed  capable  of  developing  a  boiler  horse 
power  with  two  or  three  square  feet  of  heating  surface  instead  of  ten 
square  feet,  as  at  present,  and  with  high  over-all  efficiency.  (See 
Figs.  41a  and  41b.) 

Heat  Transmission  in  Boilers,  Kreisinger  and  Ray:  Power  and  Engr.,  June  29, 
1909,  p.  1144;  U.  S.  Geological  Survey,  Bulletin  Journ.  West  Soc.  Engrs.,  Sept. 
18,  1907;  Am.  Inst.  Elect.  Engrs.,  Dec.  13,  1907. 

Heat  Transfer  and  Future  Boiler  Practice  :  A.  H.  Allen,  Power  and  Engr.,  Sept. 
21,  1909,  p.  482;  Engng.,  Lond,  Feb.  19,  1908. 

The  Heat  of  Fuels  and  Furnace  Efficiency :  W.  D.  Ennis,  Power  and  Engr., T  July 
14,  1908,  p.  50. 

A  Study  in  Heat  Transmission  (The  Transmission  of  Heat  to  Water  in  Tubes  as 
Affected  by  the  Velocity  of  the  Water),  J.  K.  Clement  and  C.  M.  Garland,  Univ. 
of  111.  Bulletin  No.  40,  Sept.  27,  1909. 

70.  Heating  Surface.  —  All  parts  of  the  boiler  shell,  flues,  or  tubes 
which  are  covered  by  water  and  exposed  to  hot  gases  constitute  the 
heating  surface.  Any  surface  having  steam  on  one  side  and  exposed 
to  hot  gases  on  the  other  is  superheating  surface.  According  to  the 
recommendations  of  the  American  Society  of  Mechanical  Engineers, 
the  side  next  to  the  gases  is  to  be  used  in  measuring  the  extent  of  the 
heating  surface.  Thus  measurements  are  made  of  the  inside  area  of  fire 
tubes  and  the  outside  area  of  water  tubes.  The  heating  surface  in  a 
boiler  under  average  conditions  of  good  practice  is  most  efficient  when 
the  heated  gases  leave  the  uptake  at  a  temperature  of  100  to  200 
degrees  F.  above  that  of  the  steam.  Each  square  foot  of  heating  surface 
is  capable  of  transmitting  a  certain  amount  of  heat,  depending  upon  the 
conductivity  of  the  material,  the  character  of  the  surface,  the  temperature 
difference  between  the  gas  and  the  water,  the  location  and  arrangement  of 
the  tubes,  the  density  of  the  gas,  the  velocity  of  the  gas,  and  the  time  allowed 
for  transmission  of  the  heat.  It  is  customary  to  assume  a  uniform  heat 


BOILERS  93 

transmission  for  the  entire  surface.  Thus  with  most  boilers  it  is  found 
that  the  best  results  are  obtained  with  an  evaporation  of  from  3  to  3.5 
pounds  of  water  from  and  at  212  degrees  F.  per  square  foot  of  heating 
surface,  which  is  equivalent  to  allowing  12  to  10  square  feet  per  boiler 
horse  power.  By  increasing  the  quantity  of  heat  the  evaporation  may 
be  increased,  but  at  the  expense  of  efficiency,  since  a  smaller  percentage 
of  the  heat  is  utilized.  For  example,  an  evaporation  as  high  as  20 
pounds  per  square  foot  per  hour  has  been  effected  in  torpedo-boat  prac- 
tice, and  12  pounds  per  square  foot  per  hour  is  not  unusual  in  locomotive 
work,  but  such  performances  are  invariably  obtained  at  the  expense  of 
economy.  The  selection  of  the  proper  proportion  of  heating  surface  to 
the  evaporation  required  is  evidently  a  very  important  matter.  For 
maximum  economy  under  average  conditions  of  operation,  practice 
allows  a  proportion  of  1  square  foot  to  every  3.5  pounds  of  water  to  be 
evaporated  from  and  at  212  degrees  F.  Where  economy  must  be  sacri- 
ficed to  capacity,  as  in  locomotive  practice,  a  much  higher  evaporation 
is  allowed. 

The  maximum  evaporation  is  limited  by  the  amount  of  coal  which  can 
be  burned  upon  the  grate.  It  the  draft  is  sufficient,  a  good  boiler  can 
develop  a  horse  power  upon  0.75  to  0.5  of  the  surface  recommended. 
In  the  very  latest  large  central  stations  the  gas  passages  and  grate 
surface  are  proportioned  so  that  the  boiler  may  be  operated  at  100% 
above  standard  rating  with  high  over-all  efficiency. 

The  following  table  shows  approximately  the  result  which  may 
be  expected  with  different  rates  of  evaporation. 


POUNDS  WATER  EVAPORATED  FROM  AND  AT  212  DEGREES  F.  PER  SQUARE 
FOOT  OF  HEATING  SURFACE  PER  HOUR. 


2 

2.5 

3 

3.5 

4 

5 

6 

8 

10 

12 

PROBABLE  RELATIVE  ECONOMY. 

100 

100 

100 

100 

99 

98 

95 

90 

85 

80 

Efficiency  of  Boiler  Heating  Surface:  Trans.  A.S.M.E.,  18-328,  19-571.  Kent,  Steam 
Boiler  Economy  (John  Wiley  &  Son),  Chapter  IX.  The  Nature  of  True  Boiler  Effi- 
ciency: Jour.  West.  Soc.  Engrs.,  Sept.  18,  1907.  Heat  Transference  through  Heating 
Surface:  Engineering,  77-1. 

71.  The  Horse  Power  of  a  Boiler.  —  A  boiler  horse  power  is  equivalent 
to  the  evaporation  of  34.5  pounds  of  water  per  hour  from  a  temperature 
of  212  degrees  F.  to  steam  at  atmospheric  pressure.  This  corresponds 


94  STEAM  POWER  PLANT  ENGINEERING 

to  33,305  B.T.U.  per  hour.*  Since  the  power  from  steam  is  developed 
in  the  engine  and  the  boiler  itself  does  no  work,  the  above  measure  of 
capacity  is  merely  conventional.  Thus  one  boiler  horse  power  will 
furnish  sufficient  steam  to  develop  about  three  actual  horse  power  in 
the  best  compound  condensing  engine,  but  only  one-half  horse  power 
in  a  small  non-condensing  engine.  Boilers  should  be  purchased  on  the 
basis  of  heating  surface  and  not  on  the  horse  power  rating,  since  one 
bidder  may  offer  a  boiler  with  say  5  square  feet  of  heating  surface  per 
horse  power  and  another  with  10  square  feet,  both  being  capable  of 
the  required  evaporation,  but  the  one  with  a  small  heating  surface 
(which  will,  of  course,  be  the  cheaper  boiler)  will  do  so  only  at  an 
increased  cost  of  fuel.  Manufacturers  ordinarily  rate  their  boilers  on 
the  basis  of  10  to  12  square  feet  of  heating  surface  per  horse  power,  and 
the  power  assigned  is  called  the  builder's  rating.  As  this  practice  is  not 
uniform,  bids  and  contracts  should  always  specify  the  amount  of  heating 
surface  to  be  furnished.  According  to  the  recommendations  of  the 
American  Society  of  Mechanical  Engineers,  "  A  boiler  rated  at  any 
stated  capacity  should  develop  that  capacity  when  using  the  best  coal 
ordinarily  sold  in  the  market  where  the  boiler  is  located,  when  fired  by 
an  ordinary  fireman,  without  forcing  the  fires,  while  exhibiting  good 
economy.  And  further,  the  boiler  should  develop  at  least  one-third 
more  than  stated  capacity  when  using  the  same  fuel  and  operated  by 
the  same  fireman,  the  full  draft  being  employed  and  the  fires  being 
crowded;  the  available  draft  at  the  damper,  unless  otherwise  under- 
stood, being  not  less  than  one-half  inch  water  column. 

In  determining  the  boiler  horse  power  required  for  a  given  engine 
horse  power  it  is  convenient  to  estimate  the  steam  consumption  of 
the  engine  under  actual  conditions  and  then  ascertain  the  equivalent 
evaporation  from  and  at  212  degrees  F.  For  example,  assume  a  single 
non-condensing  engine  developing  20  horse  power  to  use  50  pounds  of 
steam  per  horse  power  hour,  or  1000  pounds  steam  per  hour;  steam  pres- 
sure, 80  pounds  per  square  inch;  feed-water  temperature,  120  degrees  F. 
Required  the  boiler  horse  power  necessary  to  furnish  this  quantity  of 
steam. 

From  equation  (11),  the  factor  of  evaporation  is 

-,       l-t  +  32      1185.3-120  +  32 
970.4  970.4 

One  thousand  pounds  of  steam  under  the  given  conditions  are  there- 

*  With  the  new  value  of  r  =  970.4  in  place  of  965.7  this  figure  becomes 
33,478.8. 


BOILERS 


95 


fore  equivalent  to  1000   X    1.131   =    1131    pounds  from   and   at   212 
degrees  F. 

The  boiler  horse  power  necessary  to  furnish  steam  for  the  20-horse- 
power  engine  will  be 

1131 
Boiler  horse  power  =  =-:-=-  =  32.8. 

o4.O 

Example  :  A  15,000  kilowatt  steam  turbine  and  auxiliaries  require 
14.7  pounds  of  steam  per  kilowatt-hour  at  rated  load;  steam  pressure 
200  pounds  per  square  inch  gauge;  superheat  150  degrees  F.;  feed-water 
temperature,  179  degrees  F. 

Required  the  boiler  horse  power  necessary  to  furnish  this  quantity 
of  steam. 

The  heat  furnished  to  the  turbine  and  auxiliaries  per  kilowatt-hour  is 

w  l/i  +  Opts  -0-32)}=  14.7  {1199.2  +  0.57  X  150  -(179  -32)} 
=  16,724  B.T.U. 

15,000  X  16,724 
Boiler  horse  power  =  -  '     -  =  7500  (approx.). 


Table  12  gives  the  required  hourly  evaporation  per  boiler  horse  power 
at  various  feed  temperatures  and  steam  pressures. 

The  following  table  shows  approximately  the  relation  between  boiler 
horse  power  and  heating  surface  for  different  ratios  of  evaporation: 


EVAPORATION   FROM    AND  AT   212  DEGREES   F.  PER  SQUARE  FOOT 

PER  HOUR. 


2 

2.5 

3.0 

3.5 

4 

5 

6 

7 

8 

9 

10 

SQUARE  FEET   HEATING  SURFACE   REQUIRED  PER   HORSE   POWER. 

17.3 

13.8 

11.5 

9.8 

8.6 

6.8 

5.8 

>.9 

4.3 

3.8 

3.5 

Builders  of  return  tubular  and  vertical  fire-tube  boilers  allow  11  to  12 
square  feet  of  heating  surface  per  horse  power;  water-tube  boilers  are 
rated  at  10  square  feet  per  horse  power,  and  Scotch-marine  boilers  at 
8  square  feet  per  horse  power. 

72.  Grate  Surface. — The  amount  of  fuel  which  can  be  burned  per  hour 
limits  the  amount  of  water  evaporated  per  unit  of  time  and  depends 


96 


STEAM  POWER  PLANT  ENGINEERING 


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BOILERS  97 

upon  the  extent  and  nature  of  the  grate  surface,  the  character  of  the 
fuel,  and  the  draft.  A  liberal  allowance  of  grate  surface  is  usually 
desirable,  particularly  when  the  boilers  are  to  be  forced,  since  too  small 
a  grate  increases  the  labor  of  handling  and  cleaning  fires  and  results 
in  poor  economy. 

With  good  coal  low  in  ash  approximately  equal  results  may  be 
obtained  with  large  grate  surface  and  light  draft  and  with  small  grate 
and  strong  draft,  the  amount  of  coal  burned  per  hour  being  the  same 
in  both  cases.  Bituminous  coal  low  in  ash  gives  best  results  with 
high  rates  of  combustion,  provided  the  ratio  of  grate  surface  to  heating 
surface  is  properly  proportioned.  Coals  high  in  ash  require  a  compara- 
tively large  grate  surface,  particularly  if  the  ash  is  easily  fusible,  tending 
to  choke  the  grate.  Where  a  strong  draft  is  available  a  smaller  grate 
may  be  used  than  with  moderate  draft,  as  a  thicker  bed  of  fuel  can  be 
carried.  The  relation  between  draft  and  rate  of  combustion  for  various 
sizes  of  coals  is  shown  in  Fig.  116,  paragraph  127. 

A  number  of  boiler  tests  made  by  Barrus  ("  Boiler  Tests  ")  showed  that 
the  best  economy  with  anthracite  coal  hand  fired  was  obtained  with  an 
average  ratio  of  grate  surface  to  heating  surface  of  1  to  36  and  at  a  rate 
of  combustion  of  approximately  12  pounds  of  coal  per  square  foot  of 
grate  surface  per  hour.  In  these  tests  a  variation  in  grate  and  heating 
surface  ratio  of  1  to  36  up  to  1  to  46  gave  practically  no  difference  in 
economy.  With  bituminous  coal  the  tests  showed  that  an  average 
ratio  of  1  to  45  gave  the  best  results  and  at  a  rate  of  combustion  of 
24  pounds  of  coal  per  square  foot  of  grate  surface  per  hour. 

Tests  made  by  Christie  (Trans.  A.S.M.E.,  19-330)  gave  an  average 
combustion  of  13  pounds  of  anthracite  per  square  foot  of  grate  per  hour 
for  maximum  efficiency  and  24  pounds  of  bituminous. 

Current  central  station  practice  gives  normal  rates  of  combustion 
approximately  as  follows  (Ibs.  per  sq.  ft.  per  hr.) : 

Anthracite 15-20  Eastern  bituminous 20-24 

Semi-bituminous 18-22  Western  bituminous 30-35 

Table  13  gives  the  relation  between  heating  and  grate  surface  in  a 
number  of  recent  boiler  installations  using  different  kinds  of  coal,  and  is 
illustrative  of  current  practice. 

In  proportioning  the  grate  surface  for  a  proposed  installation  the 
principal  factor  considered  is  the  character  of  the  fuel,  a  study  being 
made  of  the  various  fuels  available,  and  the  one  selected  which  gives 
the  highest  evaporation  per  dollar.  The  latter  data  may  usually  be 
obtained  from  records  of  plants  using  the  same  grade  of  fuel  and  grates 
similar  to  those  intended  for  the  proposed  plant. 


98 


STEAM  POWER  PLANT  ENGINEERING 


TABLE  13. 

RATIO  OF  HEATING  SURFACE  TO  GRATE  SURFACE    IN  RECENT  BOILER 

INSTALLATIONS. 


Nature  of  Plants. 

No.  of 
Plants. 

Type  of 
Boiler. 

Type  of 
Grate. 

Height  of 
Chimney. 

Character  of 
Fuel. 

Ratio  of 
Heating 
to  Grate 
Surface. 

Central  stations 
Do.  . 

10 

8 

Hor  water 
tube. 

.  do 

Chain  
Roney 

200  feet  and 
over. 

do 

111.  screen- 
ings, 15  to 
20%  ash. 
Bituminous. 

65 
60 

Do 

6 

do 

Murphy 

do 

do 

60 

Do 

9 

do 

Miscel's 

do 

Anthracite 

40 

Manufacturing 
plants 
Office  buildings 

Central  station* 

20 
6 
1 

Return 
tubular. 
...do  

Babcock 
&  Wilcox. 

Hand  fired 

Shaking 
grates. 
Roney  

150-175 
Over  200  .  . 
Over  200  .  . 

Anthracite. 
Bituminous. 
Bituminous. 

35 

48 
31 

*  Two  stokers,  one  at  front  and  one  at  rear  of  setting.     (Power,  Jan.  7,  1908,  p.  25.) 

73.  Boiler  and  Furnace  Efficiency.  —  The  efficiency  of  the  boiler, 
including  the  grate,  is  expressed  by  the  ratio  between  the  heat  absorbed 
by  the  boiler  per  pound  of  dry  coal  fired  and  the  calorific  value  of  one 
pound  of  dry  coal.  The  efficiency  of  the  boiler  alone  is  taken  as  the 
ratio  between  the  heat  absorbed  per  pound  of  combustible  burned 
on  the  grate  and  the  calorific  value  of  one  pound  of  combustible.  The 
combustible  burned  on  the  grate  is  equal  to  the  coal  as  fired  minus 
moisture  and  the  total  refuse  in  the  ash  pit.  The  calculation  of  these 
efficiencies  is  illustrated  by  the  following  example: 

ANALYSIS    OF    COAL. 

Per  Cent. 

Moisture 8 

Ash 12 

Combustible _80 

100 

Pounds. 

Water  evaporated  from  and  at  212°  F.  per  pound  of  coal  as  fired  . . .        8.281 

Per  Cent. 

Total  refuse  in  ash  pit  16 

Percentage  of  ash  in  refuse 13 

Combustible  in  ash 3 

B.T.U. 

Heating  value  per  pound  of  coal  as  fired 11,680 

Heating  value  per  pound  of  dry  coal  =  11,680  -5-  0.92   12,696 

Heating  value  per  pound  of  combustible  =  11,680  -T-  0.80 14,600 

8.281  •*-  0.92  =  9.001  =  equivalent  evaporation  per  pound  of  dry  coal. 
9.001  X  965.7  =  8,692  =  heat  absorbed  per  pound  of  dry  coal. 


BOILERS  99 


8  692 
Efficiency  of  boiler  and  grate  =      '         =  68.49  per  cent. 


Combustible  burned  on  grate  =  100  -  (8  +  16)  =  76  per  cent. 
8.281  -T-  0.76  =  10.896  =  equivalent    evaporation    per    pound    of    com- 
bustible burned  on  the  grate. 
10.896  X  965.7  =  10,522  =  heat  absorbed  per  pound  of  combustible. 

Efficiency  of  boiler  =       '         =  72.07  per  cent. 

The  efficiency  of  the  grate  alone  might  be  expressed 

Efficiency  of  boiler  and  grate 

Efficiency  of  grate  = T^/E  • r  u  -i — ' 

Efficiency  of  boiler 

which  is  equivalent  to 

T^  .  Combustible  actually  burned 

Efficiency  of  grate  =  ~ r — pri — F^ ' 

Combustible  fired 

the  numerator  being  the  coal  fired  less  moisture  and  the  refuse  from 
the  ash  pit,  and  the  denominator  the  coal  fired  less  moisture  and  the 
ash  as  determined  from  the  proximate  analysis. 

The  efficiency  of  combustion  is  sometimes  expressed  in  terms  of  the 
difference  in  temperature  between  fuel  bed  and  flue  gas: 

rp    rp 

Efficiency  of  furnace  =  •=*• =^  '  (14) 

If  —   1  a 

in  which 

Tf  =  temperature  of  the  furnace. 

Tc  =  temperature  of  the  flue  gas. 

Ta  =  temperature  of  the  air. 

The  efficiency  of  the  furnace  or  combustion  may  also  be  stated 
(R.  S.  Hale,  Trans.  A.S.M.E.,  20-769): 

S  +  F 

Efficiency  of  furnace  =   — ^ — »  (15) 

ri 

in  which 

S  =  B.T.U.  absorbed  by  the  boiler  per  pound  of  dry  coal. 

F  =  B.T.U.  lost  in  flue  gases  per  pound  of  dry  coal. 

H  =  Calorific  value  of  1  pound  of  dry  coal. 

The  heat  absorbed  by  the  boiler  expressed  in  percentage  of  the  heat 
available  has  been  given  the  name  true  boiler  efficiency  by  the  U.  S. 
Geological  Survey  and  may  be  expressed 

fjl  fTJ 

True  boiler  efficiency  =  •=£ — ~^>  (15a) 

•*•  f       J-  s 

in  which 

Ta  =  temperature  of  the  steam  (saturated);  other  notation  as 

in  (14). 

74.  Boiler  Performances.  —  Table  14  is  compiled  from  a  number  of 
tests  of  different  types  of  boilers  with  various  types  of  grates  and 


100 


STEAM  POWER   PLANT  ENGINEERING 
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102 


STEAM  POWER  PLANT  ENGINEERING 


TABLE  14a. 

PRINCIPAL  DATA  AND  RESULTS  OF  TESTS  ON   BOILER  NO.  6,  UNIT  NO.  10, 

FISK   ST.   STATION.     COMMONWEALTH  EDISON  CO.,  CHICAGO. 

(B.  &  W.  Boiler,  "  Standard  "  Setting.) 

Water-heating  Surface,  5000  Sq.  Ft.     Superheating  Surface,  914  Sq.  Ft. 
Chain  Grate  Surface,  90  Sq.  Ft. 


Test 
No. 

Date, 
1908. 

Horse 
Power. 

Eff'y, 
PerCent. 

H.P.  per 
Sq.  Ft. 
Grate. 

Heat 
Lost  in 
Refuse, 
PerCent. 

Total 
Heating 
Surface 
per  H.P. 

Super- 
heat of 
Steam, 
Deg.  F. 

Dry  Coal 

per  Sq.  Ft. 
G.  S. 
per  Hour. 

2 

Mar.    9 

873 

67.4 

9.70 

2.8 

6.76 

197 

41.2 

4 

10 

873 

69.0 

9.52 

2.8 

6.89 

195 

39.1 

6 

11 

852 

67.3 

9.47 

2.8 

6.93 

189 

38.9 

8 

16 

836 

65.3 

9.29 

6.4 

7.06 

174 

39.5 

10 

17 

870 

68.8 

9.67 

5.0 

6.78 

180 

39.3 

14 

19 

920 

66.2 

10.22 

9.2 

6.42 

187 

43.7 

16 

23 

900 

69.5 

10.00 

4.0 

6.56 

181 

40.5 

18 

24 

916 

69.1 

10.18 

5.5 

6.44 

190 

41.6 

20 

26 

912 

69.2 

10.13 

4.4 

6.48 

179 

41.2 

22 

27 

906 

67.7 

10.07 

4.1 

6.52 

194 

42.5 

24 

30 

925 

69.8 

10.28 

2.8 

6.38 

179 

41.6 

26 

31 

894 

69.4 

9.93 

5.2 

6.60 

170 

40.6 

28" 

Apr.    1 

922 

71.2 

10.24 

3.6 

6.40 

169 

40.4 

30 

2 

923 

71.5 

10.26 

4.6 

6.40 

173 

40.5 

32 

7 

914 

70.0 

10.20 

4.5 

6.46 

175 

40.9 

34 

8 

939 

73.8 

10.4 

3.8 

6.28 

181 

40.4 

36 

10 

911 

70.9 

10.1 

3.0 

6.48 

185 

40.2 

38 

11 

967 

70.1 

10.7 

3.0 

6.11 

192 

42.6 

40 

13 

995 

67.8 

11.1 

3.4 

5.93 

211 

43.6 

42 

14 

887 

66.8 

9.9 

4.5 

6.65 

202 

40.8 

44 

27 

880 

69.5 

9.8 

5.5 

6.72 

169 

39.7 

48 

29 

927 

71.5 

10.3 

3.3 

6.37 

171 

40.8 

50 

30 

899 

70.3 

10.0 

4.2 

6.57 

171 

39.6 

52 

May    6 

886 

69.4 

9.8 

5.3 

6.67 

171 

38.2 

54 

7 

900 

69.1 

10.0 

4.8 

6.56 

171 

39.2 

56 

8 

967 

71.9 

10.7 

4.8 

6.10 

164 

40.1 

58 

11 

902 

70.5 

10.0 

3.3 

6.55 

163 

39.6 

60 

13 

875 

70.7 

9.7 

3.8 

6.74 

147 

38.3 

64 

14 

1102 

72.0 

12.2 

4.8 

5.35 

180 

43.2 

BOILERS 


103 


TABLE   14a. 

PRINCIPAL  DATA  AND  RESULTS  OF  TESTS  ON   BOILER  NO.  6,  UNIT  NO.  10, 

FISK  ST.   STATION.     COMMONWEALTH  EDISON  CO.,   CHICAGO. 

(B.  &  W.  Boiler,  "Standard"  Setting.) 

Water-heating  Surface,  5000  Sq.  Ft.     Superheating  Surface,  914  Sq.  Ft. 
Chain  Grate  Surface,  90  Sq.  Ft. 


Draft 

B.T.U. 
per  Pound 
Dry  Coal. 

Ash  in 
Dry  Coal, 
Per  Cent. 

Ash  in 
Refuse, 
Per  Cent. 

Uptake 
Temp. 
Deg.  F. 

C02, 
Per  Cent. 

Heat  Lost 
up  Stack 
(Dry  Gas), 
Per  Cent. 

Over 
Fire. 

In 
Uptake. 

.87 

1.34 

11,634 

18.46 

82.33 

466 

6.9 

.78 

.25 

11,759 

16.81 

81.36 

461 

6.7 

.83 

.25 

12,039 

16.08 

80.03 

463 

7.7 

is.'e 

.94 

.34 

11,993 

15.91 

67.42 

477 

7.6 

16.8 

.84 

.24 

11,909 

15.71 

71.32 

475 

7.9 

16.2 

.99 

.41 

11,768 

16.04 

63.78 

479 

8.5 

15.4 

.77 

.17 

11,846 

16.68 

79.04 

483 

9.1 

14.0 

.81 

.25 

11,800 

16.39 

71.98 

484 

8.3 

15.8 

.77 

.21 

11,846 

15.51 

78.53 

486 

9.0 

14.5 

.78 

.22 

11,659 

17.59 

80.58 

494 

9.2 

14.6 

.68 

.28 

11,800 

16.22 

82.97 

487 

8.8 

15.1 

.70 

.24 

11,752 

16.18 

76.84 

484 

8.8 

15.1 

.62 

.21 

11,862 

15.38 

82.99 

480 

9.2 

14.1 

.58 

.40 

11,800 

16.02 

78.37 

480 

9.1 

14.4 

.73 

.24 

11,815 

16.84 

77.84 

494 

9.0 

14.7 

.72 

.25 

11,659 

18.06 

82.27 

504 

8.9 

15.3 

.65 

.13 

11,831 

17.15 

86.92 

493 

9.7 

13.4 

.70 

.24 

12,002 

16.05 

84.39 

502 

9.0 

15.1 

.71 

.23 

12,469 

14.87 

82.14 

522 

9.7 

13.3 

.63 

.09 

12,049 

15.17 

78.12 

500 

9.5 

13.3 

.71 

.26 

11,801 

15.75 

77.21 

470 

8.3 

15.7 

.68 

.23 

11,769 

18.59 

84.04 

472 

8.7 

14.2 

.66 

.27 

11,955 

16.11 

79.30 

473 

7.9 

16.1 

.62 

.20 

12,360 

13.63 

74.59 

476 

8.8 

14.5 

.66 

.31 

12,298 

13.62 

75.19 

480 

9.0 

14.4 

.66 

1.29 

12,423 

13.37 

75.61 

474 

9.4 

13.3 

.92 

1.18 

11,956 

17.45 

83.24 

451 

9.2 

12.5 

.76 

0.98 

11,971 

17.45 

80.99 

443 

10.0 

11.2 

.68 

1.15 

13,126 

10.24 

70.90 

487 

10.4 

12.1 

104  STEAM  POWER  PLANT  ENGINEERING 

characters  of  fuel.  Although  some  of  the  tests  show  a  combined 
efficiency  of  boiler  and  grate  as  high  as  85  per  cent  (Engr.,  Lond.,  March 
21,  1902,  p.  286),  such  a  performance  cannot  be  expected  for  continuous 
operation  under  the  average  conditions  of  practice.  In  pumping 
stations  or  in  plants  where  there  are  no  peak  loads  and  the  boiler  may 
be  operated  under  a  practically  constant  set  of  conditions  a  continuous 
efficiency  of  75  per  cent  has  been  realized  with  coal  as  fuel  and  80  per 
cent  with  crude  oil,  though  these  figures  are  exceptional.  In  very  large 
central  stations,  with  the  usual  loads  in  the  morning  and  evening,  an 
average  efficiency  throughout  the  year  of  65  per  cent  is  possible,  though 
a  good  figure  is  not  far  from  60  per  cent.  In  large  isolated  stations  with 
variable  loads  good  practice  gives  an  average  of  60  per  cent.  Small 
stations  though  showing  an  efficiency  as  high  as  75  per  cent  at  times 
seldom  average  50  per  cent  for  the  year.  The  usual  discrepancy  between 
efficiency  as  determined  by  special  tests  and  everyday  operation  is  due 
to  the  fact  that  the  efficiency  test  is  usually  conducted  under  ideal  con- 
ditions :  the  boiler  surfaces  are  cleaned,  the  rate  of  combustion  carefully 
adjusted  for  maximum  economy,  and  special  attention  given  to  the  firing, 
whereas  in  actual  practice  these  refinements  are  seldom  attempted. 
Much  depends  upon  the  efficiency  of  the  boiler-room  staff,  the  character 
of  furnace  and  fuel,  draft,  and  the  load  factor.  From  the  commercial 
standpoint  the  performance  is  best  expressed  in  terms  of  the  "  cost 
to  evaporate  1000  pounds  of  water  from  and  at  212,"  or  the  "  pounds 
of  water  evaporated  per  $1  of  coal."  Table  15  gives  the  results  of  a 
number  of  tests,  made  at  the  Armour  Glue  Works,  Chicago,  111.,  show- 
ing the  cost  of  evaporating  water  with  different  grades  of  Illinois  coal. 
The  results  were  obtained  from  hand-fired  Stirling  boilers. 

Bailer  Room  Economies:  Am.  Elecn.,  Oct.,  1901,  p.  506,  Sept.,  1905,  p.  472;  Cas- 
sier's  Mag.,  March,  1906,  p.  373;  Elec.  World,  March  4,  1905;  Engr.  U.S.,  May  1,  1905, 
p.  304,  Jan.  1,  1907;  Engr.,  June  4,  1907,  p.  758;  Eng.  Rec.,  June  27,  p.  685;  Elecn., 
Lond.,  Aug.  5,  1904;  Power,  Aug.,  1905,  p.  484;  Eng.  Mag.,  Oct.,  1901,  March,  1903. 

Care  and  Management  of  Boilers :  Engr.  U.S.,  March  1,  1902,  p.  142,  Feb.  15, 1904, 
July  15,  1904,  Jan.  1,  1907;  Engineering,  Feb.  18,  1898,  p.  211,  July  15,  1898,  p.  84; 
Eng.  Mag.,  Feb.,  1901,  p.  877,  Oct.,  1901,  p.  91,  March,  1903,  p.  896;  Power,  Sept., 
1904,  p.  467,  May,  1905,  p.  267,  Dec.,  1905,  p.  742,  Sept.,  1906,  p.  550;  Am.  Elecn., 
Feb.,  April,  Sept.,  1904;  Mech.  Eng.,  July  25,  1903;  Engr.,  Lond.,  April  15,  1904; 
Elec.  Rev.,  July  13,  1907. 

75.  Effect  .of  Capacity  on  Efficiency.  —  In  general,  as  the  horse 
power  of  a  boiler  increases  above  normal  capacity  the  over-all  efficiency 
will  decrease,  due  to  the  fact  that  the  furnace  and  gas  passages  are 
ordinarily  proportioned  to  effect  an  evaporation  of  about  3.5  pounds  of 
water  from  and  at  212  degrees  F.  per  square  foot  of  heating  surface  per 
hour  at  rated  load,  the  temperature  of  the  escaping  gases  being  from 


BOILERS 


1-05 


150  to  200  degrees  above  that  of  the  steam.  To  increase  the  rate  of 
evaporation  more  coal  must  be  burned  per  unit  of  time  and  consequently 
a  larger  volume  of  gas  is  generated.  The  larger  the  volume  of  gas  the 
higher  will  be  its  velocity,  which  finally  reaches  a  point  where  heating 
surface  is  insufficient  in  extent  to  absorb  the  extra  heat  and  as  a  con- 
sequence the  flue  gas  escapes  at  a  higher  temperature.,  resulting  in 
lower  boiler  and  furnace  efficiency.  With  properly  proportioned  grate, 
furnace  and  gas  passages  a  boiler  may  be  operated  at  100%  above  stand- 
ard rating  with  little  or  no  decrease  in  over-all  efficiency.  Fig.  40 
shows  a  case  in  which  the  efficiency  decreased  with  the  increase  in 
capacity,  and  Fig.  41  a  illustrates  increased  efficiency  for  the  higher 
rates  of  driving.  These  curves  are  of  value  simply  as  illustrations  of 
the  behavior  in  specific  cases,  and  are  not  applicable  to  all  types  of 
boilers. 

TABLE  15. 

RESULTS  OF  COAL  TESTS  AT  ARMOUR   GLUE  WORKS,   CHICAGO,   AUG.    17,    1905. 


Date  of  Test. 

Name  and  Kind  of  Coal. 

Railroad  Car 
Number. 

Cost 

per 
Ton 
Deliv- 
ered. 

Cost  to 
Evaporate 
1000 
Pounds  of 

Water. 

Pounds 
Water 
Evapo- 
rated per 
$1.00  of 
Coal. 

March  5,  1905  .  . 

Williamson     County 

C.C.C.&St.  L. 

$1.90 

$0.1531 

6,532 

Coal  Co.'s,  mine  run 

No.  26368 

March  3,  1905  .  . 

Harden  &  Hafer,  mine 

S.  I.  No.  5735 

1.70 

0.1231 

8,123 

run 

June  14,  1905  .  . 

Crerar-Clinch  &  Co., 

I.C.  No.  88362 

1.50 

0.1293 

7,734 

2"  screenings. 

June  15,  1905  .  . 

....do  

I.C.  No.  88362 

1.50 

0.1218 

8,210 

June  16,  1905  .  . 

....do  

I.  C.  No.  88362 

1.50 

0.1175 

8,511 

June  17,  1905  .  . 

Brackett    Coal    and 

C.  &E.  I.,  No. 

1.65 

0.122 

8,197 

Coke  Co.,  lump. 

8891. 

June  19    1905 

do 

C.  &  E.  I.  No. 

1.65 

0.1212 

8,251 

5002  ' 

June  20,  1905  .  . 

....do  

C.  &  E.  I.  No. 

1.65 

0.1352 

7,396 

5002 

July  1,  1905.  .  .  . 

Kellyville    Coal    Co., 

C.  &  E.  I.  No. 

1.595 

0.1355 

7,380 

mine  run. 

10030. 

July  6,  1905.  .  .  . 

Brackett  C.  &  C.  Co., 

C.  &  E.  I.  No. 

1.65 

0.1236 

8,091 

Keeler  mine  run. 

12367 

July  28,  1905.  .  . 

Kellyville    Coal    Co., 

C.  &E.  I.  No. 

1.50 

0.1285 

7,782 

washed  pea. 

6211. 

July  29,  1905.  .  . 

....do  

C.  &E.  I.  No. 

1.50 

0.119 

8,403 

6211 

Aug.  5,  1905  .  .  . 

Bering  Coal  Co.,  mine 

C.  &  E.  I.  No. 

1.575 

0.125 

8,000 

run. 

25125 

Aug.  7,  1905  .  .  . 

Bering  Coal  Co.,  Sulli- 

E. &T.  H.  No. 

1.40 

0.11 

9,091 

van  Co.,  screenings. 

5132. 

Aug.  8,  1905  .  .  . 

Consolidated  Indiana 

E.  &T.  H.  No. 

1.35 

0.105 

9,524 

Coal   Co.,    Sullivan 

3239 

Co.,  screenings. 

Aug.  9,  1905  .  .  . 

Screenings  

E.  &T.  H.  No. 

1.30 

0.0973 

10,277 

6534 

Aug.  11,  1905  .. 

Ziegler,  screenings  .  .  . 

I.  C.  No.  81184 

1.50 

0.1047 

9,551 

*  See,  "  The  Nature  of  True  Boiler  Efficiency,"  Jour.  Wes.  Soc.  Engrs.,  Oct., 
1907,  p.  677. 


106 


STEAM  POWER  PLANT  ENGINEERING 


In  nearly  all  stations  the  boilers  must  have  sufficient  overload 
capacity  to  take  care  of  peak  loads  or  to  allow  some  of  the  boilers  to  be 
shut  down  for  cleaning  or  repairs,  since  the  installation  of  sufficient 


70 


Jour. 


0.2 


Vv-.S.E.,Fert 


17  1904. 


0.45 


0.5 


CO 

w 
fi 

fc£h 

5-5S> 

«a 


.01 

0 


ra 
ou 


4.5  £ 


* 

W  0 

Lb.  Wate 


0.25  0.3  0.35  0.4 

Draft  over  Fire  in  Inches  of  Water/ 

FIG.  40.   Influence  of  Draft  on  the  Efficiency  and  Capacity  of  a  350-Horse-power  Babcock 
and  Wilcox  Boiler  with  Chain  Grate. 


Lb.  Water  per  Lb.  Combustlble(212°F) 

*°  °  £•  * 

X 

.  —  - 

"-^> 

-^ 

r 

< 

5^, 

Sx 

"X 

S 

s^ 

s 

x 

N 

s 

X 

X 

X 

™ 

23456 

Lb.  Water  Evaporated  per  Sqi  Ft.  of  Heating  Surface  per  Hour 

FIG.  41.    Effect  of  Rate  of  driving  on  Economy  of  a  150-Horse-power  Stirling  Boiler, 

Hand  Fired. 

rated  boiler  capacity  would  be  expensive  and  in  many  instances  pro- 
hibitive in  cost.  In  small  stations,  however,  too  large  a  boiler  capacity 
frequently  is  to  be  preferred  to  an  overloaded  installation,  since  the 


BOILERS 


107 


•      • 

• 

• 
• 

• 
^** 

• 
•  ^* 

x^ 

• 

^ 

• 

i*^ 
• 

• 

•    • 

• 
..' 

•  ^> 

• 

• 

• 

• 

• 
• 

• 

••• 

X 

P* 

• 

•  • 

•  • 

• 

• 

X 

• 

• 

• 

• 
•  •• 

• 

• 
• 

• 

/. 

•" 

• 
• 

• 

• 

• 

/ 

• 

'/ 

500  H.P. 
5000  Sq.  Ft.  V 
940  „       -,    g 
Standard  Cha 
90  Sq.  Ft.  Gra 
Various  Grac 
Street  Station,  Co 
Chi 

B.  &A 
fater 
uperl 
in  Gn 
teSui 
esof 
mmo 
cago  ] 

^..Boiler 
Heating  Surface 
leating         •• 
ite  and  Setting 
face 
Coal 
Q  Wealth-Edison 

:n. 



/ 

/ 

/ 

Fiski 

Co. 

/ 

7 

• 

05 


55 


50 


45 


400 


500 


GOO 


700  800  900 

Boiler  Horse  Power 


1000 


1100  1200 


FIG.  41a.     Relation  Between  Efficiency  and  Capacity,  500  H.P.  Boiler,  Fisk  Street 
Station,  Commonwealth  Edison  Co.,  Chicago. 


600  H.P.   B.&  W.   BOILER 
6008  SQ.FT.  HEATING  SURFACE 


EQUIPPED  WITH  RONEY  STOKERS 

AT  THE  59TH  ST.  STATION 

OF  THE  INTERBOROUGH  RAPID 

TRANSIT  CO.,  N.Y. 


Coal  14250  B.Tj.U.per  Pound 


i 


840  920  1000  1080 

Boiler  Horse  Power 


1160 


1240         1320 


FIG.  41b.     Effect  of  Rate  of  Driving  on  the  Efficiency  of  a  600  H.P.  B.  &  W.  Boiler. 


108 


STEAM  POWER  PLANT  ENGINEERING 


1300 


1200 


1100 


;iooo 


o 
£  900 


800 


700 


/ 

/ 

X 

x1 

A 

/ 

B/ 

C 

X 

X 

/ 

/ 

/ 

/ 

/ 

/ 

f 

/ 

/ 

/ 

</ 

/ 

\/ 

DOUBLE  STOKER 
Front,  100  Sq.Ft. 

/ 

••  :' 

X 

^o 

^ 

^> 

A  ^tatfcjDraft  over  En 
B        ••         ••      in  Heat 
—  C  Velocity  Draft  in  Fl 
Immediately  outsid 
toiler  Setting 

:e 
Pass 
ue 
eof 

0 

/ 

o 

x 

Xj 

x 

/ 

B/ 

/ 

/ 

x 

A/ 

f 

/ 

A 

/ 

SINGS' 

L^ri 

OKE 

100  S 

I  Ft. 

/ 

0        . 

/, 

^r 

/ 

/ 

// 

/ 

600  H.P.   B.&  W.  BO 
6008  SQ.FT.  HEATINGS 
RONEY  STOKERS 
59THST.  STATION,  INTER 
RAPID  TRANSIT  CO. 

ILER 
UR.FACE 

BOROUGH 
,N.Y. 

/ 

y/ 

Y 

/ 

0* 

t 

0  0.1  0.2  0.3  0.4  0.5  0.6  0.7  0.8 

Draft  in.  Inches  of  Water 

FIG.  41c.     Influence  of  Draft  on  the  Capacity  of  a  600  H.P.  B.  &  W.  Boiler. 


TABLE    15a. 

BOILER    PERFORMANCE. 

Pounds  of  Water  Evaporated  per  Hour  from  and  at  212  Deg.  F.  per  pound  of  Fuel. 


Calorific  Value 
of  Fuel, 
B.T.U. 
per  Pound. 

Boiler  and  Furnace  Efficiency. 

40 

45 

50 

55 

60 

65 

70 

75 

80 

85 

7,500 

3.09 

3.48 

3.86 

4.25 

4.64 

5.02 

5.41 

5.80 

6.18 

6.57 

8,000 

3.30 

3.71 

4.12 

4.55 

4.95 

5.36 

5.77 

6.18 

6.60 

7.01 

8,500 

3.51 

3.94 

4.38 

4.81 

5.26 

5.70 

6.14 

6.57 

7.01 

7.45 

9,000 

3.71 

4.18 

4.64 

5.10 

5.56 

6.04 

6.50 

6.96 

7.42 

7.90 

9,500 

3.92 

4.41 

4.90 

5.39 

5.88 

6.47 

6.86 

7.35 

7.85 

8.33 

10,000 

4.12 

4.64 

5.16 

5.66 

6.19 

6.70 

7.21 

7.74 

8.25 

8.76 

10,500 

4.31 

4.86 

5.40 

5.94 

6.48 

7.01 

7.55 

8.10 

8.64 

9.17 

;   11,000 

4.52 

5.09 

5.65 

6.22 

6.79 

7.35 

7.91 

8.48 

9.05 

9.61 

11,500 

4.74 

5.31 

5.91 

6.50 

7.10 

7.69 

8.28 

8.86 

9.45 

10.0 

12,000 

4.94 

5.55 

6.16 

6.78 

7.40 

8.01 

8.64 

9.25 

9.86 

10.5 

:  12,500 

5.14 

5.78 

6.42 

7.06 

7.70 

8.35 

9.00 

9.64 

10.3 

11.0 

13,000 

5.35 

6.01 

6.69 

7.35 

8.01 

8.69 

9.35 

10.0 

10.7 

11.4 

13,500 

5.56 

6.25 

6.95 

7.65 

8.34 

9.03 

9.72 

10.4 

11.1 

11.8 

14,000 

5.75 

6.48 

7.20 

7.91 

8.64 

9.35 

10.1 

10.8 

11.6 

12.2 

14,500 

5.96 

6.70 

7.45 

8.20 

8.95 

9.70 

10.5 

11.2 

12.0 

12.7 

15,000 

6.18 

6.95 

7.72 

8.50 

9.26 

10.1 

11.8 

11.6 

12.4 

13.1 

BOILERS 


109 


extra  first  cost  of  the  former  may  be  less  than  the  loss  due  to  poor 
efficiency  and  depreciation  in  the  latter. 

As  far  as  forcing  is  concerned  the  fire-tube  boiler  is  as  effective  as  the 
water-tube,  more  depending  upon  the  furnace,  grate  surface,  draft  and 
character  of  fuel  than  upon  the  type  of  boiler.  All  boilers  are  subject 
to  more  or  less  priming  at  heavy  overloads,  and  the  overload  capacity 
is  often  limited  on  this  account. 

The  Forcing  Capacity  of  Fire-Tube  Boilers:  F.  W.  Dean,  Trans.  A.S.M.E.,  26-92. 
Increasing  Capacity  of  Steam  Boilers:  Kreisinger  and  Ray,  Power,  May  24,  1910. 

76.  Thickness  of  Fire.  —  For  a  given  furnace  and  boiler,  quality  and 
size  of  fuel  and  intensity  of  draft,  a  certain  depth  of  fuel  will  give  maxi- 
mum efficiency.  Too  thin  a  fire  results  in  an  excess  of  air  and  too 
thick  a  fire  in  a  deficiency,  the  economy  being  lowered  in  either  case. 
On  account  of  the  number  of  conditions  upon  which  the  proper  thick- 
ness depends,  it  can  only  be  determined  for  a  particular  case  by  actual 
test,  the  available  data  being  insufficient  for  drawing  conclusions.  The 
curves  in  Fig.  42  are  plotted  from  a  series  of  tests  made  on  a  350-horse- 


400 


300 


100 


4  5 

Thickness  of  Fire,  Inches 


FIG.  42.    Effect  of  thickness  of  Fire  on  the  Capacity  and  Efficiency  of  a  350-Horse-power 
Stirling  Boiler,  equipped  with  Chain  Grate. 


power  Stirling  boiler  equipped  with  chain  grate  at  the  power  plant  of 
the  Armour  Institute  of  Technology.  The  damper  was  left  wide  open 
throughout  the  test  and  the  speed  of  the  grate  kept  constant.  Ratio 
of  grate  to  heating  surface,  1  to  42.  Carterville  washed  coal  No.  4  was 
used  in  all  tests.  The  curves  in  Fig.  43  refer  to  the  performance  of  a 
150-horse-power  water-tube  boiler  equipped  with  chain  grate  at  the 
University  of  Illinois  Engineering  Experiment  Station  at  Urbana. 


110 


STEAM  POWER  PLANT  ENGINEERING 


The  curves  in  Fig.  44  are  plotted  from  a  series  of  tests  on  a  500-horse- 
power  Babcock  &  Wilcox  boiler  equipped  with  chain  grate  at  the  Fisk 
Street  station  of  the  Commonwealth  Edison  Company,  Chicago,  111.  In 
these  tests  the  conditions  of  operation  are  not  exactly  comparable, 
but  they  serve  to  show  the  variation  of  economy  with  thickness  of  fire 
in  each  case.  In  general,  with  natural  draft,  fine  sizes  of  coal  necessi- 
tate thin  fires,  since  they  pack  so  closely  as  to  greatly  restrict  the  draft. 
Thin  fires  require  closer  attention  to  prevent  holes  being  burned  in 


I 


I  £  7;  120 

P?    o    . 

I  §  j"° 

43  b  8° 
P 


Capacity  of  Boile  • 


s  2 

8  13  ° 

"o  "3  "§ 
§ 


6.5 


5.5 


6-iu. 


Fire 


15  20  25  30  35  40 

Dry  Coal  per  Sq.  Ft.  of  Grate  Surface  per  Hr.-Lbs. 

FIG.  43.    Effect  of  Thickness  of  Fire  on  the  Capacity  and  Efficiency  of  a  150-Horse-power 

Water-Tube  Boiler. 


spots,  and  respond  less  readily  to  sudden  demands  for  steam,  but  have 
the  advantage  of  letting  the  air  required  pass  through  the  grate,  whereas 
thick  fires  often  require  air  to  be  supplied  above  the  grate  to  insure 
complete  combustion.  Thick  fires  require  less  attention  and  hence  are 
preferred  by  firemen.  Where  sufficient  draft  is  available  thick  fires 
are  more  efficient  than  thin  ones,  as  the  air  excess  is  more  readily 
controlled. 


BOILERS 


111 


77.  Influence  of  Initial  Temperature  on  Efficiency.  —  In  general  the 
higher  the  initial  temperature  of  the  furnace  the  greater  will  be  the 
efficiency  of  the  heating  surface,  since  the  heat  transmitted  varies  almost 
directly  with  the  difference  of  temperature  between  the  water  and  the 
products  of  combustion.  If  the  heating  surface  is  properly  distributed 
so  that  the  final  temperature  of  the  escaping  gas  remains  constant,  the 
efficiency  of  the  boiler  and  furnace  will  increase  as  the  initial  temperature 
increases,  though  not  in  direct  proportion.  This  is  on  the  assumption 


1000 
*   800 

S   600 

1 

400 

200 
0 

£    80 

<a 
1     60 

^i 

*H 

-«-~ 

--< 

•—  «. 

*•«« 

S 

***-, 

s^ 

*** 

^ 

X 

^ 

<^ 

~~, 

--*, 

X 

^, 

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1 

40 

A   Boiler  14  Tubes  High 
B      "        9      «« 
Jour.  W.S.E.  11-539 

4  5  6  7  8  9  10  11  12 

Thickness  of  Fire  in  Inches 

FIG.  44.    Effect  of  Thickness  of  Fire  on  the  Capacity  and  Efficiency  of  a  500-Horse-powet 
Babcock  and  Wilcox  Boiler. 

that  the  amount  of  heat  generated  per  hour  is  the  same  throughout 
all  ranges  in  temperatures.  With  a  condition  where  the  amount  of 
heat  generated  remains  constant  and  the  initial  temperature  varies, 
the  final  temperature  of  the  escaping  gases  remains  practically  constant, 
and  in  such  cases  high  initial  temperatures  are  productive  of  high 
boiler  and  furnace  efficiencies.  In  practice  these  conditions  are  seldom 
realized  and  high  furnace  temperatures  are  not  necessarily  productive 
of  high  boiler  and  furnace  efficiencies.  Some  tests  show  a  decided  gain 
in  efficiency  with  the  higher  furnace  temperatures  ("Some  Perform- 


112  STEAM  POWER  PLANT  ENGINEERING 

ances  of  Boilers  and  Chain  Grate  Stokers,  with  Suggestions  for  Improve- 
ments/' A.  Bement,  Jour.  West.  Soc.  Engrs.,  February,  1904),  and 
others  show  little  if  any  improvement  ("  A  Review  of  the  United  States 
Geological  Survey  Fuel  Tests  under  Steam  Boilers,"  L.  P.  Breckenridge, 
Jour.  West.  Soc.  Engrs.,  June,  1907).  The  majority  of  high  efficiency 
records,  however,  are  associated  with  high  furnace  temperatures. 

78.  Cost  of  Boilers  and  Settings.  —  Figures  giving  the  cost  of  boilers, 
irrespective  of  type,  at  so  much  per  horse  power  are  misleading,  since 
the  cost  does  not  increase  in  the  same  ratio  as  the  power.     The  wide 
variation  in  cost  on  the  horse-power  basis  is  partly  due  to  the  difference 
in  rating.     For  instance,  Scotch  marine  boilers  are  ordinarily  rated  at 
8  square  feet  of  heating  surface  per  horse  power  and  return  tubular  boilers 
at  12  square  feet.     The  price  approximates  one  dollar  per  square  foot  of 
heating  surface  for  all  boilers  over  100  horse  power.     The  cost  of  water - 
tube  and  fire-tube  boilers  may  be  roughly  estimated  by  the  following 
formulas  (C.  H.  Benjamin,  Engr.  U.S.,  Nov.  15,  1902): 

(A)  Cost  in  dollars  =  500  +  9.2  X  rated  horse  power.  (16) 

(B)  Cost  in  dollars  =  500  +  8.5  X  rated  horse  power.  (17) 

(C)  Cost  in  dollars  =  100  +  6.5  X  rated  horse  power.  (18) 

(D)  Cost  in  dollars  =  100  +  5.0  X  rated  horse  power.  (19) 

(A)  Horizontal  water-tube  boilers,  125  pounds  pressure,  10  square  feet  heating 
surface  per  horse  power. 

(B)  Vertical  water-tube  boilers,  other  conditions  as  in  (A). 

(C)  Horizontal  return  tubular  boilers,  12  square  feet  heating  surface  per  horse 
power. 

(D)  Small  vertical  fire-tube  boilers. 

The  cost  of  Scotch-marine  boilers  rated  on  a  basis  of  8  square  feet 
per  horse  power  may  be  estimated  by  means  of  formula  (A). 

The  cost  of  plain  settings  may  be  roughly  approximated  as  follows: 

Horizontal  water  tube: 

Cost  =  400  +  0.8  X  rated  horse  power.  (20) 

Return  tubular: 

Cost  =  300  +  0.7  X  rated  horse  power.  (21) 

79.  Selection   of  Type.  —  Boilers   constructed   by  builders   of   good 
repute  are  usually  designed  for  safety,  durability,  and  capacity,  and 
rigid  specifications  and  inspection  of  material  and  workmanship  are 
ordinarily   not  necessary,    as   the   makers'   reputations   are   sufficient 
guarantee  of  their  worth.      Marked  departure  from  standard  designs 
must  necessarily  be  specified,  but  in  most  cases  instructions  are  limited 
to  the  extent  of  heating  and  grate  surface,  the  character  of  the  furnace, 


BOILERS  113 

and  arrangement  of  setting.  Numerous  tests  on  various  types  of  boilers 
show  practically  the  same  efficiency  provided  the  furnaces  and  boilers 
are  properly  designed,  so  that  the  relative  merits  may  be  considered 
with  reference  to  (1)  durability;  (2)  accessibility  for  repairs;  (3)  facility 
for  cleaning  and  inspection;  (4)  space  requirements;  (5)  adaptability 
to  the  type  of  furnace  and  stoker  desired;  (6)  capacity;  and  (7)  cost  of 
boiler  and  setting.  For  high  pressure,  150  pounds  per  square  inch  or 
more,  the  water -tube  or  some  form  of  internally  fired  boiler  in  which  the 
shell  plates  are  not  exposed  to  the  high  temperature  of  the  furnace  is 
considered  safer  than  the  horizontal  tubular  boiler  because  the  shell 
plates  and  the  seams  of  the  latter  must  be  of  considerable  thickness  in 
large  units,  and  being  exposed  to  the  hottest  part  of  the  fire  are  likely 
to  give  trouble,  especially  if  the  water  contains  scale  or  sediment -form- 
ing elements.  Return  tubular  and  stationary  locomotive  boilers  are 
seldom  made  in  sizes  over  200  horse  power  and  hence  are  not  to  be 
considered  for  large  units.  For  sizes  over  150  horse  power  where  over- 
head room  is  limited  the  return  tubular  boiler  is  most  commonly 
installed,  unless  high  pressure  is  essential,  in  which  case  the  internally 
fired  Scotch-marine  boiler  is  peculiarly  adaptable.  The  water -tube 
boiler  is  usually  employed  in  large  central  stations  for  high -pressure 
units  of  300  to  500  horse  power.  The  particular  type  of  water -tube 
boiler  is  to  some  extent  a  matter  of  personal  taste  on  the  part  of  the 
engineer.  For  small  powers  and  for  intermittent  operation,  small 
vertical  or  horizontal  fire-box  boilers  have  the  advantage  of  low  first 
cost.  The  small  air  leakage  and  radiation  losses  give  internally  fired 
boilers  an  advantage  over  externally  fired  fire-tube  or  water -tube  types, 
but  this  is  partly  offset  by  the  greater  extent  of  regenerative  sur- 
face in  the  setting  of  the  latter.*  Internally  fired  boilers  are  more  ex- 
pensive than  the  externally  fired,  though  the  extra  cost  of  setting  and 
foundation  in  the  latter  may  bring  the  total  cost  of  the  entire  equip- 
ment to  practically  the  same  figure.  The  design  and  installation  of  the 
boilers  and  furnaces  should  be  left  at  the  outset  to  a  capable  engineer. 

Makers  usually  request  the  following  information  from  intending  pur- 
chasers: i.   Steam  pressure  desired. 

2.  The  quantity  of  steam  demanded. 

3.  The  kind  of  fuel  to  be  burned. 

4.  The  type  of  furnace  or  stoker. 

5.  The  nature  and  intensity  of  draft. 

6.  Nature  of  setting. 

7.  Probable  temperature  of  the  feed  water. 

*  At  the  power  plant  of  the  Cosmopolitan  Electric  Co.,  Chicago,  the  brick  settings 
of  the  boilers  (500  h.p.  B.  &  W.)  are  completely  encased  with  riveted  boiler  plates. 


114 


STEAM  POWER  PLANT  ENGINEERING 


The  complete  specifications  for  a  return  tubular  boiler  are  given  in 
Chapter  XIX. 

Choice  of  Boilers  for  Steam  Power  Plants  :  Am.  Elecn.,  June,  1900,  p.  261,  Dec., 
1905,  p.  633,  May,  1901,  p.  217,  June,  1903,  p.  256;  Engr.,  Lond.,  Sept.,  1898,  p.  232, 
May  29,  1903,  p.  555;  St.  Ry.  Rev.,  Oct.,  1901,  p.  789;  Elec.  Rev.,  Lond.,  June  16, 
1899,  p.  973;  St.  Ry.  Jour.,  Oct.  5,  1901,  p.  453,  Oct.,  1905,  p.  731;  Mech.  Eng., 
May  16,  1903;  Min.  Rept.,  Feb.  21,  1907. 

80.  Grates.  —  Grates  may  be  divided  into  three  general  classes, 
namely,  stationary,  rocking,  and  traveling  grates.  The  latter  are 
treated  in  Chapter  IV.  Stationary  grates  are  generally  made  of  cast- 
iron  sections  in  a  variety  of  shapes  as  illustrated  in  Fig.  45.  The 
bars  are  ordinarily  from  3  inches  to  4  inches  deep  at  the  center  (this 
makes  them  strong  enough  to  carry  the  load  caused  by  the  weight  of 
the  fuel  without  sagging  even  when  the  top  is  red  hot),  f  inch  wide  at 
the  top,  and  taper  to  f  inch  at  the  bottom  to  enable  the  ashes  to  drop 
clear.  The  width  of  the  air  space  is  determined  by  the  size  of  the  fuel 
to  be  used,  the  average  proportions  being  given  in  Table  16. 

TABLE   16. 

AIR    SPACES   AND    THICKNESS    OF    GRATE    BARS. 


Size  and  Kind  of  Coal. 

Width  of  Air 
Spaces. 

Thickness  of 
Grate  Bars. 

Screenings 

(Inch) 
i 

* 

1 

f 

itof 
*to  | 

(Inc 
1 

] 

h) 
f 

r 

Anthracite  — 
Average            ....       ...       

Buckwheat  .        

Pea  or  nut  

Stove 

Eeff 

Broken                                           

Lump                        

Bituminous  average 

Wood- 
Slabs                                             

Sawdust                             

Shavings                           

The  "  Tupper  "  and  "  Herringbone  "  grate  bars  are  stiffer  and  less 
likely  to  warp  than  the  common  form,  but  are  not  so  readily  sliced  and 
therefore  not  so  convenient  with  coal  that  clinkers  badly.  Sawdust 
or  pin-hole  grates  are  used  in  burning  sawdust,  tan  bark,  and  very  small 
sizes  of  coal.  Grates  are  often  set  horizontally  and  the  bars  are  held 
in  place  simply  by  their  own  weight,  but  long  grates  are  best  placed 
sloping  toward  the  rear  to  facilitate  firing.  The  front  of  the  grate  when 


BOILERS 


115 


designed  for  bituminous  coal  is  often  made  solid,  this  portion  being 
called  the  "  dead  plate."  It  serves  to  hold  the  green  fuel  until  the 
hydrocarbons  have  been  distilled  off,  when  the  charge  is  pushed  back 
on  the  open  grate  at  the  time  of  next  firing.  The  length  of  a  single  bar 
or  casting  should  not  exceed  three  feet.  The  length  of  grate  may  be 


COMMON    BAR 


TUPPER 


HERRINGBONE 


••••••••••••••••  ••••!   /I    ,  •  ••• 

*J •••••••••••  \  VjJ  i*s^ 


SAW-  DUST 

FIG.  45.    Types  of  Grate  Bars. 

made  of  two  or  three  bars  and  should  not  exceed  6  feet  with  bituminous 
coal,  as  this  is  the  greatest  length  of  fire  that  can  be  readily  worked  by 
a  stoker.  With  buckwheat  anthracite  furnaces  12  feet  in  depth  are  not 
unusual,  as  anthracite  fires  require  no  slicing. 

The  disadvantage  of  using  stationary  grates  is  that  the  fire  is  not 
easily  cleaned.  Unless  the  air  spaces  are  kept  free  of  clinkers  and 
ashes,  combustion  is  hindered  and  the  fire  rendered  sluggish.  Frequent 
cleaning,  however,  is  wasteful  of  fuel  and  reduces  the  furnace  efficiency 
by  letting  in  a  large  excess  of  air  every  time  the  fire  door  is  opened. 

81.  Rocking  Grates.  —  Shaking  grates  have  the  advantage  of  per- 
mitting stoking  without  opening  the  fire  door  and  require  less  manual 
labor  than  stationary  grates.  There  is  a  great  variety  of  sectional 
shaking  grates  on  the  market  and  some  of  them  are  made  self-dumping. 
One  of  the  best  known  types  is  illustrated  in  Fig.  46.  Each  row  or 
section  of  grate  bars  is  divided  into  a  front  and  a  rear  series  by  twin 
stub  levers  and  connecting  rods.  An  operating  handle  is  adapted  to 
manipulate  either  one  or  both  of  the  levers  in  such  a  manner  that  the 
front  and  rear  series  may  operate  separately  or  together.  The  shaking 
movement  causes  no  increase  in  the  size  of  the  openings  and  hence  pre- 
vents the  waste  of  fine  fuel.  Ordinarily  the  width  of  the  grate  is  made 


116 


STEAM  POWER  PLANT  ENGINEERING 


equal  to  two  or  more  rows  of  grate  bars  so  that  the  live  fire  may  be 
shoved  sidewise  from  one  row  to  the  other  when  cleaning.     A  depth  of 


FIG!  46.   A  Typical  Rocking  Grate. 

fire  of  from  6  to  10  inches  is  carried  according  to  the  nature  of  the  fuel 
and  the  available  draft. 

Grate  Bars:  Engr.  U.S.,  Nov.  1,  1906,  p.  728,  Jan.  1,  1907,  p.  68;  Am.  Elecn., 
Jan.,  1904,  p.  269. 

82.  Blow-Offs.  —  Boilers  must  be  provided  with  blow-off  pipes  for 
draining  off  the  water  and  for  discharging  sediment-  and  scale-forming 
material.  The  "  bottom  blow  "  is  ordinarily  an  extra  heavy  pipe  of 
suitable  diameter  connected  to  the  lowes't  part  of  the  boiler  and  fitted 


FIG.  47.    Horizontal  Blow-off  Connection 
to  Head. 


FIG.  48.   Vertical  Blow-off  Connection 
to  Shell. 


with  a  valve  or  cock,  or  both.  (See  paragraph  349.)  Fig.  47  shows  an 
arrangement  of  horizontal  blow-off  connected  to  the  head  of  a  return 
tubular  boiler  and  Fig.  48  a  vertical  blow-off  connected  to  the  shell. 


BOILERS 


117 


The  latter  is  the  better  arrangement.  The  blow-off  pipe  where  it 
passes  through  the  back  connection  is  covered  with  magnesia,  asbestos, 
or  fire  brick.  When  exposed  to  the  action  of  extremely  hot  furnace 
gases  as  in  forced  draft  installations,  the  arrangement  illustrated  in  Fig. 
49  is  sometimes  used  to  prevent  the  pipe  from  burning  out.  When  the 
blow-off  cock  is  shut  and  the  valve  on  the  vertical  branch  open  there 
is  a  continuous  circulation  of  water.  Where  boilers  are  arranged  in 
batteries,  the  battery  may  have  a  common  outlet  for  the  blow-off 


_  __   WATER   LEVEL. 

3 

5 

8 

0 

H 

St 

H 

BLOW    C 

)FF 

FIG.  49.    Blow-off  Connection  with 
Circulating  Pipes. 


FIG.  50.    Blow-off  Tank  and 
Connections. 


pipes,  as  illustrated  in  Fig.  388.  Usually  the  blow-off  pipes  may  dis- 
charge into  the  open  air,  but  this  is  not  permissible  in  large  cities,  nor 
is  it  lawful  to  blow  directly  into  the  sewer.  In  this  case  the  water 
and  sediment  may  be  discharged  into  a  blow-off  tank  and  permitted  to 
cool  before  delivery  to  the  sewer,  as  illustrated  in  Fig.  50. 


WATER  LEVEL 


FIG.  51.    Surface  Blow-off. 


"  Surface  blows  "  are  often  installed  to  remove  scum,  grease,  and  float- 
ing or  suspended  particles  of  dirt.  The  bell -mouthed  shape  shown  in 
Fig.  51  permits  the  skimmer  to  accommodate  itself  to  varying  water 


118 


STEAM  POWER  PLANT  ENGINEERING 


level,  and  it  is  sometimes  provided  with  a  float  and  with  a  flexible  joint, 
Fig.  52. 

83.   Damper    Regulators.  —  For    maximum    furnace    efficiency    the 
draft  must  be  regulated  to  burn  just  enough  fuel  to  supply  the  steam 


FIG.  52.    Buckeye  Skimmer. 

required.  Where  forced  draft  is  employed  this  is  done  by  regulating 
the  speed  of  the  blower.  With  natural  draft  it  is  the  usual  practice  to 
regulate  the  draft  by  means  of  dampers  placed  in  the  uptake,  and  in 
order  that  the  regulation  may  be  effective  it  should  be  automatic. 
Automatic  dampers  are  economical  and  useful  and  are  particularly 
desirable  in  plants  where  the  demand  for  steam  fluctuates  rapidly. 
There  are  several  successful  types  on  the  market,  some  operated  by 
water  pressure  and  others  by  direct  boiler  pressure.  Fig.  53  illus- 


FIG.  53.    Kitts  Hydraulic  Damper  Regulator. 

trates  such  a  mechanism.  Full  boiler  pressure  acting  at  all  times  on  a 
diaphragm  A  raises  or  lowers  a  weight  W  attached  to  arm  D  according 
as  the  steam  pressure  increases  or  decreases.  Arm  D  actuates  a  small 
valve  V  which  controls  the  supply  of  water  to  chamber  B.  A  diaphragm 
in  chamber  B  raises  and  lowers  the  damper  as  the  water  pressure  varies, 
a  drop  of  0.5  pound  being  sufficient  to  open  the  damper  to  its  maximum. 


BOILERS 


119 


The  steam  diaphragm  has  a  movement  of  only  0.01  inch  and  the 
water  diaphragm  0.5  inch.  When  properly  adjusted  and  given  proper 
attention  automatic  dampers  work  in  a  very  satisfactory  manner. 

Fig.  54  shows  a  section  through  the  Tilden  damper  regulator,  illus- 
trating the  principles  of  the  steam-actuated  type.     The  device  is  con- 


STEAM 


FIG.  54.   Tilden  Steam  Actuated 
Damper  Regulator. 


WATER 

FIG.  55.    Simple  Water  Column. 


nected  directly  to  the  boiler  by  pipe  A.  The  pressure  on  piston  B  is 
balanced  by  spring  C  under  normal  conditions  of  operation.  Any 
variation  from  the  normal  will  cause  the  rod  R  to  move  up  or  down,  so 
that  the  dampers  are  opened  or  closed  in  proportion  to  the  change  in 
pressure.  The  chamber  N  is  separate  from  chamber  M,  so  that  steam 
or  water  cannot  come  in  contact  with  the  spring.  Piston  D  acts  as  a 
guide  only. 

Damper  Regulators:  Engr.  U.S.,  Jan.  1,  1907,  p.  58;  Elec.  Wld.,  May  2,  1908. 

84.  Water  Gauge.  —  The  water  level  in  a  boiler  is  usually  indicated 
either  by  a  gauge  glass,  by  try  cocks,  or  both,  connected  directly  to  the 
boiler  as  in  Fig.  1,  or  to  a  water  column  or  combination  as  in  Fig.  55. 


120 


STEAM  POWER  PLANT  ENGINEERING 


Each  gauge-glass  connection  should  be  fitted  with  a  stop  valve  which 
may  be  closed  in  case  the  tube  breaks.  In  large  boilers  these  valves, 
usually  of  the  quick-closing  type,  are  conveniently  operated  from  the 
boiler-room  level  by  means  of  a  chain  attached  to  the  valve  stem. 
Self-closing  automatic  valves  are  frequently  employed,  one  type  being 
illustrated  in  Fig.  56.  If  the  glass  breaks  the  outrush  of  steam  forces  the 
ball  against  a  conical  seat  and  shuts  off  the  supply.  When  a  new  glass 
is  inserted  the  ball  is  forced  back  by  slowly  screwing  in  the  valve  stem. 
Hinged  valves  mechanically  operated  from  without  are  considered 
more  reliable  than  ball  valves,  as  scale  is  less  likely  to  render  them 
inoperative. 


FIG.  56.    Water  Guage  with  Self-closing 
Valve. 


FIG.  57.    Combined  Water  Column 
and  High  and  Low- Water  Alarm. 


Try  cocks  or  gauge  cocks  are  set  at  points  above  and  below  the  desired 
water  level,  preferably  connected  directly  to  the  boiler  shell  but  some- 
times to  a  water  column  as  in  Fig.  55.  The  water  level  is  ascertained 
by  opening  the  cocks  in  succession. 

The  objection  to  the  latter  arrangement  is  that  accident  to  or  a 
stoppage  of  the  piping  renders  both  gauge  glass  and  try  cocks  useless. 
Water  columns  should  be  blown  out  once  a  day  and  the  gauge  cocks 
opened  to  see  that  the  height  of  the  water  indicated  tallies  with  that 
shown  by  the  glass.  Some  engineers  prefer  two  separate  columns  to 
each  boiler  and  no  cocks,  others  rely  solely  upon  cocks. 

The  water  column  shown  in  Fig.  57  has  an  alarm  whistle,  controlled 
by  two  floats,  which  gives  a  high  and  low-water  alarm.  Numerous 
devices  of  this  class  are  on  the  market  but  they  are  usually  regarded  as 


BOILERS 


121 


unreliable  and  most  engineers  are  content  to  depend  upon  water  gauge 
and  try  cocks. 

Water  Gauges  and  Columns:  Mach.,  Sept.,  1905,  p.  31;  Power,  Aug.,  1905,  p.  483; 
Am.  Elecn.,  July,  1904,  p.  359;  Engr.  U.S.,  Jan.  1,  1907,  p.  58. 

85.   Fusible  or  Safety  Plugs.  —  Fusible  or  safety  plugs  as  illustrated 
in  Fig.  58  are  brass  plugs  provided  with  a  fusible  metal  core.     They 


TYi>ES 


FIG.  58.   Types  of  Fusible  Plugs. 

are  inserted  in  the  shell  or  tubes  at  the  lowest  permissible  water  line. 
When  covered  by  water  the  heat  is  conducted  away  sufficiently  fast  to 
keep  the  temperature  below  the  fusing  point,  but  when  uncovered  the 
low  conductivity  of  the  steam  prevents  the  rapid  withdrawal  of  heat, 
whereupon  the  alloy  melts  and  the  blast  of  escaping  steam  gives  warning. 
The  melting  point  of  fusible  metals  being  sometimes  uncertain,  plugs 
occasionally  blow  out  without  apparent  cause  and  at  other  times  fail  to 
act  when  shell  is  overheated.  Fusible  plugs  are  required  by  law  in 
many  cities. 

86.  Mechanical  Tube  Cleaners.  —  Although  purifying  plants,  boiler 
compounds,  and  the  like  are  preventive  of  scale  formation  to  a  great 
extent,  experience  shows  that  the  most  satisfactory  method  is  to  use 
mechanical  tube  cleaners  for  cutting  or  breaking  the  scale.  The  prin- 
ciples of  construction  of  these  devices  vary  widely  according  to  the  types 
of  boilers  in  which  they  are  used,  and  depend  upon  the  nature  of  the 


C  D  F  t, 

FIG.  59.    Mechanical  Tube  Cleaner,  —  Hammer  Type. 

duty  which  they  must  perform.    They  may  be  conveniently  divided  into 
two  classes: 

1 .  Those  which  loosen  the  scale  by  a  series  of  rapid  hammer  blows, 
Fig.  59. 

2.  Those  which  cut  out  the  scale  by  a  revolving  tool,  Fig.  60. 


122 


STEAM  POWER  PLANT  ENGINEERING 


The  hammer  device  is  applicable  to  either  the  water-  or  fire-tube  type 
of  boiler,  but  the  revolving  cutter  is  applicable  to  the  water-tube  only. 
Steam,  compressed  air,  or  water  under  pressure  may  be  used  as  the  motive 
power,  though  the  latter  is  the  most  convenient  and  satisfactory. 

Referring  to  Fig.  59,  the  hammer  head  J  is  given  a  rapid  motion, 
which  may  reach  1,500  vibrations  per  minute,  and  subjects  the  tube 
to  repeated  shocks,  thereby  cracking  the  brittle  scale  and  jarring  it 
loose  from  the  water  surface  of  the  tube.  The  cleaner  head  is  -attached 
to  a  flexible  pipe  of  sufficient  length  to  enable  it  to  be  pushed  from  one 
end  to  the  other.  Even  if  carefully  manipulated  the  hammer  is  apt  to 
injure  the  tube  by  swaging  it  to  a  larger  diameter',  producing  crystalliza- 
tion in  the  metal  and  causing  leaks  where  the  tubes  are  expanded  into 
the  sheets,  hence  its  use  is  not  to  be  recommended. 

Hydraulic  turbine  cutters  are  made  in  many  designs,  one  of  which 
is  shown  in  Fig.  60.  The  cylindrical  casing  D  contains  a  hydraulic 

JC          A  T 


FIG.  60.    Mechanical  Tube  Cleaner,  —  Turbine  Type. 

turbine  consisting  of  a  fixed  guide  plate  which  directs  the  water  at  the 
proper  angles  upon  the  vanes  of  the  turbine  wheel  T.  The  cutters  C 
revolve  at  high  speed  and  chip  the  scale  into  small  pieces.  The  stream  of 
water  flowing  from  the  turbine  envelops  the  cutters,  keeps  their  edges 
cool,  and  washes  away  the  scale  as  fast  as  it  is  detached.  Different 
styles  of  cutter  wheels  are  furnished  with  each  cleaner  so  as  to  adapt 
the  device  to  all  kinds  of  scale  formations.  In  well-managed  plants 
scale  is  not  permitted  to  deposit  to  a  thickness  greater  than  ^  to  ^  of 
an  inch. 

The  soot  and  cinders  which  accumulate  on  the  inside  surface  of  fire- 
tube  boilers  are  removed  by  mechanical  scrapers,  brushes,  or  steam-jet 
blowers.  (For  a  description  of  these  devices  see  American  Electrician, 
April,  1904,  p.  576.) 

The  tubes  of  a  water-tube  boiler  are  cleaned  externally  by  means  of 
a  steam  jet. 

Boiler  Accessories :  Am.  Elecn.,  April,  1903,  p.  194,  Feb.,  1905,  p.  67,  June,  1904, 
p.  269,  July,  1904,  p.  339;  Am.  Mach.,  April  21,  1901,  p.  518;  Engr.  U.S.,  Jan.,  1907, 
p.  56. 

Boiler  Arches :  Boiler  Maker,  Aug.,  1907. 


BOILERS  ,  123 

Blow-off  Connections:  Locomotive,  Oct.,  1906;  Elec.  World,  Nov.  2,  1907;  Nat. 
Engr.,  June,  1904;  Eng.  Rec.,  May  9,  1908. 

Bracing:  Boiler  Maker,  April,  1905;  Mach.,  Sept.,  1903,  p.  18,  Oct.,  1903,  p.  83; 
Power,  Jan.,  1903,  p.  24,  Oct.,  1905,  p.  611,  Nov.,  p.  687;  Eng.  News,  Dec.  15,  1904, 
p.  533;  Trans.  A.S.M.E.,  18-989;  Am.  Mach.,  June  3,  1897,  June  2,  1898,  p.  404; 
Engr.  U.S.,  Jan.  1,  1907,  p.  18;  Engr.,  Lond.,  April  25,  1900,  pp.  412,  419;  Prac. 
Engr.,  Jan.,  1907. 

Boiler  Cleaning  :  Am.  Elecn.,  Dec.,  1900,  April,  1904,  p.  174;  Power,  May  and  Oct., 
1905,  Aug.,  1906,  p.  465;  Locomotive,  Oct.,  1904;  Boiler  Maker,  Aug.,  1905;  Engr. 
U.S.,  Jan.  1,  1907,  p.  109. 

Boiler  Design:  Engr.  U.S.,  Jan.  15,  1902,  p.  59;  Eng.  Mag.,  May,  1904,  p.  233; 
Eng.  Rec.,  July  14,  1900,  May  18,  1901,  p.  467,  Oct.  12,  1901,  p.  347;  Power,  Oct., 
1901,  p.  14,  March,  1906,  p.  147;  St.  Ry.  Rev.,  Feb.  15,  1899,  p.  125;  West.  Elecn., 
April  20,  1901,  p.  267;  Am.  Mach.,  April  21,  1904;  Mach.,  Oct.,  1902. 

Boiler  Dimensions  :  All  Types  of  Stationary  Boilers  :  Eng.  U.S.,  Jan.  1,  1907,  p.  10, 
Aug.  1,  15,  22,  1903.  Small  Marine  Boilers :  Am.  Mach.,  Sept.  3,  1896,  p.  823. 
Tubular  Boilers :  Mach.,  Oct.,  1902,  p.  94. 

Circulation  in  Boilers:  Eng.  Rec.,  July  20,  1901;  Cassier's  Mag.,  Jan.,  1905;  Elec. 
Rev.,  Lond.,  April  4,  1902;  Engng.,  April  18,  1902;  Engr.,  Lond.,  Nov.  6,  1903;  Am. 
Mach.,  Jan.  14,  1897,  p.  40,  Sept.  20,  1900,  p.  910;  Eng.  News,  Jan.  18,  1900,  p.  40; 
Trans.  A.S.M.E.,  7-814,  9-489;  Engr.  U.S.,  Oct.  15,  1907. 

Domes:  Engr.  U.S.,  Jan.  1,  1907,  p.  27. 

Classification  of  Boilers  and  Comparison  of  Types :  Engr.  U.S.,  Jan.  1,  1907;  Min. 
Rept.,  Feb.  21,  1907. 

Furnace  and  Settings:  National  Engr.,  Sept.  and  Nov.,  1907;  Elec.  World.,  Sept. 
7,  1907;  Am.  Elecn.,  Jan.,  1902,  p.  10,  Nov.,  1903,  p.  557,  July,  1904,  p.  339;  Trans. 
A.S.M.E.,  6-118,  16-590,  19-74,  782,  20-95;  Am.  Mach.,  Aug.  18,  1898;  Engr.  U.S., 
July  15,  1905,  p.  471,  Sept.  15,  1905,  p.  622,  Aug.  1,  1906,  p.  491,  May  15,  1906,  Jan. 
1,  1907;  Power,  March  24,  1908,  p.  445,  June,  1905;  Eng.  Mag.,  July,  1897,  p.  587. 

Boiler  Inspection:  Power,  Jan.,  1906,  p.  32;  Engr.  U.S.,  Feb.  15,  1907;  Trans. 
A.S.M.E.,  4-142;  Boiler  Maker,  Nov.,  1907;  Cassier's,  Feb.,  1907. 

Riveted  Joints :  Power,  March,  1906,  p.  147,  April,  1906,  p.  227;  Engr.  U.S.,  Jan.  1, 
1907,  p.  21,  Aug.  15,  1907,  p.  784;  Trans.  A.S.M.E.,  6-120,  10-707;  Boiler  Maker, 
June,  1906,  Dec.,  1907;  Prac.  Engr.,  Dec.  13,  1907. 

Safety  Valves  :   See  paragraph  350. 

Specifications:  Power,  Dec.,  1905,  p.  728;  Nat.  Engr.,  May  15,  1903,  p.  367; 
Eng.  News,  Oct.  20,  1898,  p.  251;  Boiler  Maker,  Sept.,  1906,  p.  243. 

Thickness  of  Boiler  Plate  :  Am.  Mach.,  Jan.  16  and  Feb.  27,  1902;  Trans.  A.S.M.E., 
22-127,  15-629,  24-921;  Eng.  News,  Jan.  31,  1901,  p.  121. 

Boiler  Testing :  See  A.S.M.E.  Code  for  conducting  Standard  Boiler  Trials,  reprinted 
in  Appendix  B.  See  also  Power  and  Engr.,  Feb.  23,  1909. 

Bridge  Walls  in  Theory  and  Practice:  Power  and  Engr.,  Mar.  9,  1909,  p.  452. 


CHAPTER  IV. 

SMOKE  PREVENTION,   FURNACES,  STOKERS. 

87.  General.  —  It  is  recognized  that  bituminous  coal  can  be 
efficiently  burned  without  smoke  in  a  properly  designed  furnace  if 
proper  attention  is  given  to  stoking  and  other  factors  involved.  It  is 
nevertheless  a  common  statement  among  owners  of  power  plants  that 
it  is  cheaper  to  smoke  than  to  operate  without  smoke.  This  is 
undoubtedly  true  in  many  cases  where  smokeless  combustion  can  be 
secured  only  by  admitting  a  considerable  excess  of  air  with  a  consequent 
loss  in  economy  frequently  greater  than  that  due  to  incomplete  com- 
bustion and  smoke.  Even  under  the  most  favorable  conditions,  how- 
ever, smokeless  combustion  depends  largely  upon  skillful  manipulation 
by  an  interested  and  efficient  fireman.*  The  order  of  intelligence 
demanded  for  this  work  is  out  of  all  proportion  to  the  wages  paid.  In 
many  small  plants  —  and  these  are  usually  the  most  obstinate  smoke 
offenders  —  the  fireman  handles  as  much  as  a  ton  of  coal  per  hour  by 
hand,  besides  caring  for  the  feed  pumps  and  water  levels,  keeping  the 
boilers  clean,  and  removing  the  ash.  The  boiler  room  is  frequently 
poorly  lighted  and  poorly  ventilated.  It  is,  therefore,  not  surprising 
that  the  fireman  seldom  worries  about  the  smoke  problem.  A  better 
wage  scale  and  more  consideration  for  the  firemen  might  do  a  great  deal 
toward  abating  the  smoke  nuisance.  In  order  that  combustion  may  be 
smokeless  and  efficient,  the  volatile  gases  and  separated  free  carbon  must 
be  brought  into  intimate  contact  with  the  proper  quantity  of  air  and  main- 
tained at  a  temperature  above  the  ignition  point  until  oxidation  is  complete 
before  they  are  brought  in  contact  with  the  heat-absorbing  surfaces  of  the 
boiler.  Mere  excess  of  air  will  not  effect  smokeless  combustion,  even  if 
the  gases  and  air  are  thoroughly  mixed,  if  the  temperature  is  prema- 
turely reduced  below  that  necessary  for  combustion  by  contact  with  the 
heat-absorbing  surfaces  of  the  boiler. 

Smoke  may  be  produced,  therefore,  by 

1.  An  insufficient  amount  of  air  for  the  perfect  combustion  of  the 
volatile  gases.     This  is  primarily  a  question  of  draft. 

2.  An  imperfect  mixture  of  air  and  combustible. 

3.  A  temperature  too  low  to  permit  complete  oxidation    of    the 
volatile  combustible. 

*  See  Appendix  G.  Rules  for  firemen  using  Illinois  and  Indiana  coal  in  hand- 
fired  furnaces. 

124 


SMOKE  PREVENTION,  FURNACES,  STOKERS  125 

Smoke-preventing  devices  may  be  divided  into  two  classes: 

(1)  Those  which  are  an  integral  part  of  the  boiler  and  setting,  such 
as  mechanical  stokers,  Dutch  ovens,  down-draft  furnaces  and  fire-tile 
combustion  chambers  incorporated  with  the  regular  setting. 

(2)  Those  which  may  be  conveniently  attached  to  plants  already  in 
operation  without  material  modification  of  the  furnace,  such  as  steam 
jets  and  other  means  of  mixing  air  and  combustible  gases,  admission  of 
air  through  the  bridge  wall  or  side  wall,  and  mechanical  draft. 

88.  Mechanical  Stokers.  —  Uniform  evolution  of  the  volatile  gases  of 
the  fuel  is  the  essential  requisite  for  smokeless  combustion,  and  it  is 
for  this  reason  that  mechanical  stokers  as  a  class  are  more  effective  in 
preventing  smoke  than  any  apparatus  accompanied  by  intermittent 
firing.  Stokers  which  feed  irregularly  have  the  effect  of  hand-fired 
furnaces,  and  it  is  necessary  not  only  to  employ  some  powerful  auxiliary 
mixing  device  but  also  to  furnish  at  times  an  extra  supply  to  take  care 
of  the  enormous  volume  of  volatile  gas  evolved  after  a  fresh  charge  of 
fuel  is  added. 

Carefully  adjusted  automatic  stokers  owe  their  high  efficiency  to 
(1)  uniformitij  of  feed;  (2)  proper  proportion  of  air  and  combustible;  (3) 
absence  of  excessive  air  dilution,  as  when  the  fire  doors  are  opened  in 
connection  with  hand  firing;  and  (4)  self-cleaning  grates. 

Daily  records  are  essential  with  any  type  of  stoker  or  hand  firing  if 
efficient  results  are  expected,  as  only  by  frequent  observation  is  it  possible 
to  determine  the  proper  adjustment  of  air  supply,  depth  of  fire,  rate  'of 
feed,  and  the  like.  Control  of  air  supply  is  almost  as  important  as  the 
upkeep  and  effective  operation.  In  the  best  firing  practice  the  right 
amount  of  air,  depth  of  fire,  and  rate  of  feed  must  be  worked  out  by  the 
engineer. 

Stokers  are  often  condemned  by  owners  as  inefficient  and  inferior  to 
hand  stoking  because  no  particular  attention  had  been  paid  to  them 
beyond  filling  the  hopper  with  coal.  They  should  be  operated  in  strict 
accordance  with  the  principles  of  design. 

In  plants  of  2000  horse  power  or  over  the  installation  of  mechanical 
stokers  and  coal  conveyors  effects  a  considerable  saving  of  labor  and 
can  usually  be  relied  upon  to  solve  the  smoke  problem  if  reasonable 
attention  is  given  to  their  operation.  In  smaller  plants  interest  on 
the  investment  and  other  considerations  may  make  hand  firing  more 
economical,  although  many  plants  of  capacities  as  small  as  200  horse 
power  are  giving  satisfaction,  particularly  in  places  where  a  poor  grade 
of  fuel  is  used  and  smoke  ordinances  are  rigidly  enforced.  A  stoker 
of  the  self-cleaning,  slow-running  type  requires  much  less  attention 
than  the  hand-fired  furnace.  With  hand  firing  one  fireman  can  effi- 


126  STEAM  POWER  PLANT  ENGINEERING 

ciently  attend  to  the  water,  coal,  and  ashes  of  about  200  horse  power 
or  handle  coal  for  say  500  horse  power,  whereas  with  good  automatic 
stokers  he  can  readily  take  care  of  2000  horse  power  or  of  4000  horse 
power  with  chain  grates  equipped  with  overhead  bunkers  and  down 
spouts. 

The  best  stokers  are  those  which  are  least  complicated  and  simplest 
in  operation.  A  cheap  stoker  is  a  poor  investment,  since  the  cost  of 
repairs  and  shut  downs  will  usually  amount  to  more  than  the  saving 
in  price. 

The  following  outline  gives  a  classification  of  a  few  of  the  best  known 
American  mechanical  stokers: 

Front  Feed. 

Chain  Grates:  Step  Grates: 

Babcock  &  Wilcox,  3*         Roney, 
Green,  Wilkinson, 

McKenzie,  Acme, 

Playford.  McClave. 

Side  Feed.  Under  Feed. 
Step  Grates:  Jones, 

Murphy,  American, 

Detroit,  Taylor, 

Mode.  Guckett. 

Down  Draft.  Sprinkler. 

Hawley  Little  Giant, 

Swift, 

Powdered  Fuel.  Vulcan. 

See  paragraphs  34-46. 


Mechanical  Stokers:  Power,  March,  1906,  p.  189,  Aug.,  1905,  p.  487,  March,  1903, 
p.  112 ;  Jour.  West.  Soc.  Engrs.,  Feb.,  1903,  p.  44 ;  Eng.  Soc.  West  Penn.,  April,  1903, 
p.  169 ;  Eng.  News,  March 26, 1903,  p.  272 ;  Eng.  News,  35-226 ;  Eng.  Mag.,  July,  1902, 
p.  528;  Engr.  U.S.,  Jan.  1,  1907,  p.  83,  Aug.  15,  1906,  p.  440,  July  2,  1906,  p.  437, 
Feb.  1, 1904,  p.  114,  April  1, 1903,  p.  262 ;  Elec.  Engr., Lond.,  33-977 ;  Cassier's,  Sept., 
1906,  p.  469;  Trans.  A.S.M.E.,  17-278,  558;  Am.  Engr.,  July,  1905,  p.  281,  July, 
1904,  p.  284;  Am.  Elecn.,  July,  1904,  p.  329,  1902,  p.  489,  14-18,  12-411,  263. 

89.  Chain  Grates.  —  The  chain  grate,  Fig.  61,  is  one  of  the  most 
popular  forms  of  automatic  stokers.  It  embodies  a  moving  endless 
chain  of  grate  bars  mounted  on  a  frame  with  provision  for  the  con- 
tinuous and  uniform  feeding  of  coal  into  the  furnace,  the  fuel  and  the 
grate  moving  together.  The  operations  of  feeding  the  coal,  carrying 
it  through  the  progressive  stages  of  combustion,  removing  the  ashes  and 
clinkers,  and  maintaining  a  clean  grate  and  free  air  supply  are 


SMOKE  PREVENTION,  FURNACES,  STOKERS 


127 


128 


STEAM  POWER  PLANT  ENGINEERING 


FIG.  62.    Babcock  and  Wilcox  Boiler,  Chain  Grate,  Ordinary  Setting. 


FIG.  63.   Babcock  and  Wilcox  Boiler,  Chain  Grate,  Fire-tile  Roof. 


SMOKE  PREVENTION,  FURNACES,  STOKERS 


129 


cally  automatic.  The  driving  mechanism  consists  of  a  gear  train 
actuated  by  a  cast-steel  ratchet  and  pawls,  the  arms  carrying  the  latter 
being  given  a  reciprocating  motion  by  an  eccentric  mounted  on  a- line 
shaft.  The  latter  may  be  driven  by  any  type  of  engine  or  motor  and 
the  speed  of  the  grate  regulated  by  varying  the  stroke  of  the  arm  carry- 
ing the  pawls.  Fuel  is  fed  into  a  hopper  placed  at  the  front  end  of  the 
furnace  and  the  depth  of  the  fuel  regulated  by  a  guillotine  damper. 
The  entire  grate  and  driving  mechanism  are  mounted  on  a  permanent 


FIG.  64.     Section  of  Tiles  Encircling  Lower  Row  of  Tubes. 

truck  and  may  readily  be  removed  from  beneath  the  boiler.  The  front 
part  of  the  furnace  is  provided  with  a  flat  or  slightly  inclined  ignition 
arch  as  illustrated.  The  thickness  of  the  fire  and  the  speed  of  the  grate 
should  be  so  regulated  that  when  the  fuel  has  reached  the  end  of  the 
grate  it  shall  have  been  completely  consumed  and  ashes  only  will  be  dis- 
charged into  the  pit.  The  combination  of  chain  grate  with  inclined  igni- 


FIG.  65.     Tiles  Between  Lower  Row  of  Tubes,  Back  of  Encircling  Tiles. 


tion  arch,  curved  bridge  wall,  and  lower  course  of  tubes  covered  with  fire 
tiling,  as  shown  in  Fig.  61,  makes  an  excellent  smokeless  furnace,  though 
the  depth  of  fire  and  air  supply  must  be  carefully  regulated  to  prevent 
excessive  air  dilution.  With  chain-grate  stokers  there  may  be  a  con- 
siderable leakage  of  air  between  the  grate  and  bridge  wall,  through  the 
coal  in  the  hoppers,  under  the  coal  gate,  and  through  the  fire  bed  at 
the  rear  where  it  is  mostly  ash.  Various  schemes  have  been  employed 
to  prevent  leakage  at  the  end  of  the  grate  by  using  water  backs,  ash-pit 
dampers,  and  the  like.  Fig.  62  shows  the  application  of  the  chain  grate 


130  STEAM  POWER  PLANT  ENGINEERING 

to  a  Babcock  &  Wilcox  boiler  with  the  standard  arrangement  of  fire 
tiles  and  bridge  wall.  To  insure  the  best  coking  effect  with  this  ar- 
rangement the  flat  ignition  arch  is  employed.  This  reduces  the  velocity 
and  increases  the  temperature  of  the  gases  as  they  are  distilled  from 
the  green  fuel,  thereby  assisting  complete  combustion.  Under  normal 
conditions  of  operation  this  insures  practically  smokeless  combustion, 
but  is  not  always  successful  at  heavy  overloads,  since  the  length  of 
furnace  is  not  always  great  enough  to  thoroughly  mix  the  air  and 
combustible  gases  before  they  reach  the  boiler  surfaces.  A  more 
satisfactory  arrangement  is  the  modification  illustrated  in  Fig.  63.  The 
coking  arch  is  slightly  inclined  and  the  gases  are  compelled  by  means 
of  the  fire-tile  roof  to  take  a  longer  path  to  the  rear  before  crossing 
the  tubes  to  the  uptake,  which  must  necessarily  be  at  the  front  end 
of  the  boiler.  Tests  have  shown  this  arrangement  to  give  excellent 
efficiency  of  boiler  and  furnace,  with  practically  smokeless  combustion, 
though  the  cost  of  the  upkeep  of  the  tiling  is  very  high. 

Figs.  64  and  65  illustrate  the  method  employed  in  suspending  the 
fire  tiles.  Another  method  of  increasing  the  extent  of  regenerative 
fire-brick  surface  without  disturbing  the  standard  setting  is  to  cover 
the  tubes  over  the  furnace  with  a  fire  tile,  which  permits  the  gases  to 
flow  between  the  tubes.  Such  an  application  to  a  Stirling  boiler  is 
illustrated  in  Figs.  66  and  67.  The  useful  life  of  this  class  of  tiling  is 
very  short  and  this  fact  prevents  its  adoption  in  most  cases. 

Fig.  67a  and  Fig.  67b  show  recent  applications  of  chain  grates  to 
large  B.  &  W.  boilers  which  are  effecting  high  efficiency  and  heavy  over- 
load capacity  with  practically  smokeless  combustion.  See  Power  and 
The  Engineer,  May  31,  1910,  p.  981. 

Chain  grates  properly  installed  and  taken  care  of  cost  very  little  for 
maintenance.  At  the  South  Side  Elevated  Railroad  Station,  Chicago, 
where  there  are  22  B.  &  W.  chain  grates,  the  total  cost  of  repairs  for 
the  grates  alone  is  stated  to  have  been  but  $25  in  eight  years.  Flat 
fire-brick  arches  as  in  Fig.  62  have  to  be  renewed  about  once  a  year. 
The  inclined  arch  with  a  good  quality  of  fire  brick  should  last  two  years 
or  more.  The  cost  of  renewing  arches  of  this  type  approximates  $200 
for  500-horse-power  boilers  per  setting.  The  fire-tile  furnace  is  more 
costly  in  yearly  repairs  than  the  common  furnace,  but  the  increased 
efficiency  may  offset  the  extra  cost.  (A.  Bement,  Eng.  U.  S.,  June  1, 
1907,  p.  606.)  The  repairs  for  a  500-horse-power  Babcock  &  Wilcox 
boiler  setting  as  in  Fig.  63  is  approximately  $350  per  year  for  arches 
and  tiling. 

90.  Step  Grates,  Front  Feed.  —  Fig.  68  shows  the  general  arrange- 
ment of  a  Roney  stoker  and  Fig.  69  that  of  a  Wilkinson  stoker,  illus- 


SMOKE  PREVENTION,  FURNACES,  STOKERS  131 


FIG.  66.    Application  of  "Economy"  Fire  Tiles  to  Stirling  Boiler. 


SECTION  THROUGH  A-B 


1 

L 

E 

[ 

t 

[ 

[ 

E 

FIG.  67.    Method  of  Anchoring  "  Economy"  Fire  Tiles  to  Tubes. 


132 


STEAM  POWER  PLANT  ENGINEERING 


SMOKE  PREVENTION,  FURNACES,  STOKERS 


183 


134 


STEAM  POWER  PLANT  ENGINEERING 


trating  the  step-grate,  front-feed  principle.  The  Roney  stoker  consists 
of  a  hopper  for  receiving  the  coal,  a  set  of  rocking  stepped  grates  inclined 
at  a  proper  angle  from  the  horizontal,  and  a  dumping  grate  at  the  bottom 
of  the  incline  for  receiving  and  discharging  the  ash  and  clinkers.  The 
dumping  grate  is  divided  into  several  sections  for  convenience  in  han- 
dling. The  coal  is  fed  on  to  the  inclined  grate  from  the  hopper  by  a 
reciprocating  "  pusher  "  actuated  by  the  "  agitator."  The  power  is 
supplied  through  an  eccentric  operated  by  a  small  engine  or  motor. 
The  normal  feed  is  about  10  strokes  per  minute.  The  grate  bars  rock 


Details  of  Construction  of  the  Roney  Mechanical  Stoker 
FIG.  68.      Details  of  Ronev  Stoker. 


through  an  arc  of  30  degrees,  assuming  alternately  horizontal  and 
inclined  positions.  The  construction  permits  abundance  of  air  to  pass 
through  the  fuel,  with  little  or  no  possibility  of  coal  dropping  through 
the  grate.  A  coking  arch  of  fire  brick  is  sprung  across  the  furnace  as 
indicated.  This  stoker  operates  with  natural  draft  and  with  suitable 
arrangement  of  fire  tiling  effects  complete  and  efficient  combustion. 
Without  a  fire  tile  roof  construction  smokeless  combustion  is  effected 
with  difficulty,  particularly  at  heavy  loads. 

In  the  Wilkinson  stoker  the  inclined  grate  bars  are  hollow  and  are 
arranged  side  by  side,  every  alternate  bar  being  movable.     When  in 


SMOKE   PREVENTION,  FURNACES,  STOKERS 


135 


FIG.  68a.     Double  Stoker  Installation  at  the  59th  Street  Station  of  the 
Interborough  Rapid  Transit  Co.,  N.  Y. 


136 


STEAM   POWER  PLANT  ENGINEERING 


operation  there  is  a  constant  sawing  action  of  the  grate  bars,  causing 
the  fuel  to  flow  forward  and  downward.  A  small  steam  jet  with  about 
T^  inch  opening  is  introduced  into  the  end  of  each  hollow  grate  bar, 
thus  inducing  the  required  amount  of  air  for  combustion,  which 
passes  through  air  openings  approximately  J  inch  wide  by  3  inches 


THE  MECHANISM  OP  THE  WILKINSON  STOKER. 


FIG.  69.      Details  of  Wilkinson  Stoker. 


long.  These  stokers  are  driven  by  two  small  hydraulic  motors,  the 
water  being  furnished  by  a  small  pump  and  being  used  over  and  over 
again. 

91.  Step  Grates,  Side  Feed.  —  Fig.  70  shows  a  front  vertical  section 
and  Fig.  71  a  side  vertical  section  through  a  Murphy  automatic  stoker 
and  furnace.  The  apparatus  is  in  effect  a  Dutch  oven  equipped  with 
an  automatic  feeding  and  stoking  device.  Coal  is  introduced  either 
mechanically  or  by  hand  into  the  magazine  at  each  side  of  the  furnace 
and  above  the  grate  and  descends  by  gravity  upon  the  coking  plate. 
Reciprocating  stoker  boxes  push  the  coal  out  upon  the  grate  bars. 
Every  alternate  grate  bar  is  movable  and  pivoted  at  its  upper  end. 
A  rocker  bar  driven  by  a  small  motor  or  engine  causes  the  lower  ends  to 
move  up  and  down,  this  action  producing  the  required  stoking  effect. 
A  device  for  grinding  up  the  clinker  and  ash  is  provided  as  shown  at 
the  bottom  of  the  furnace.  This  is  hollow  and  is  connected  by  a  2-inch 


SMOKE  PREVENTION,  FURNACES,  STOKERS  137 


TWNSYERSL    SECTION. 
FIG.  70.   Murphy  Furnace,  Front  Section. 


............  j.. ... L. ...  I  .TT771..... I. 


EVERY  ALTERNATE  ORATE  BAR  16  MOVA 
AND  THE  INTERMEDIATE  ONES  ARE  STATIONARY 


FIG.  71.    Murphy  Furnace,  Side  Section. 


138  STEAM  POWER   PLANT  ENGINEERING 

pipe  with  the  smoke  flue,  so  that  the  cold  air  passing  through  it  prevents 
it  being  destroyed  by  the  heat.  Air  is  supplied  to  the  green  coal  through 
flues  passing  under  the  coking  plates,  and  the  speed  of  the  stoker  boxes 
and  grate  bars  can  be  regulated  to  conform  to  any  rate  of  combustion. 
On  account  of  the  large  fire-brick  combustion  chamber,  this  stoker  with 
careful  manipulationis  capable  of  practically  smokeless  combustion.  The 
power  house  of  the  Northwestern  Elevated  Railroad  Company,  Chicago, 
111.,  is  equipped  with  Murphy  furnaces,  which  are  operating  smoke- 
lessly  at  an  unusually  high  combustion  rate,  whereas  a  number  of 
other  installations  using  the  same  type  of  stoker  and  boiler  and 
burning  the  same  class  of  fuel  are  heavy  smoke  producers.  Murphy 
furnaces  are  peculiarly  adapted  to  variable  loads,  since  at  light  loads 
the  stoker  may  be  operated  with  reduced  grate  area  by  allowing  the 
bottom  of  the  grate  to  partly  fill  with  ashes. 

92.   Underfeed  Stokers.  —  Fig.  72  shows  the  general  principles  of  the 
Jones  underfeed  stoker.     It  consists  of  a  steam-actuated  ram  with  a 


FIG.  72.    Jones  Underfeed  Stoker. 

fuel  hopper  outside  the  furnace  proper  and  a  retort  or  fuel  magazine 
and  auxiliary  ram,  A,  A,  within.  Heavy  cast-iron  tuyere  blocks  for 
the  admission  of  air  are  placed  on  either  side  of  the  retort.  Fuel  is 
forced  underneath  the  fire  by  the  ram  and  its  auxiliary,  the  ram 
movement  of  the  fuel  being  backward  and  upward,  displacing  the  incan- 
descent fuel  and  ash  and  forcing  it  on  to  the  dead  plates.  There  is 
no  ash  pit,  the  ashes  being  raked  from  the  dead  plate  by  hand.  Air, 
supplied  by  a  blower,  is  admitted  through  the  openings  in  the  tuyere 
blocks.  The  latter  are  at  a  point  above  the  green  fuel  in  the  retort 
tube,  but  below  the  fire.  The  standard  size  of  retort  is  about  6  feet 
in  length  by  24  inches  in  width  and  18  inches  in  depth,  and  experience 
has  shown  that  other  sizes  are  not  necessary,  since  the  spaces  between 
retort  and  side  walls  of  the  various  furnaces  may  be  provided  for  by 
extending  the  width  of  dead  plate.  One  stoker  is  usually  installed  for 
each  furnace,  though  two  are  sometimes  required.  The  draft  from 
the  fan  and  the  number  of  strokes  of  the  ram  are  automatically  con- 


SMOKE  PREVENTION,   FURNACES,  STOKERS 


139 


trolled  by  the  steam  pressure,  although  provision  is  made  for  regulating 
either  by  hand.  Underfeed  stokers  are  adaptable  to  all  grades  and  sizes 
of  coal  and  on  account  of  the  forced  draft  are  capable  of  burning  very 
low  grades  of  fuel.  A  number  of  these  stokers  are  installed  in  the 
power  plant  of  the  First  National  Bank  Building,  Chicago,  and  are 
giving  high  efficiency  and  smokeless  combustion  with  low-grade  Illinois 
screenings.  The  cost  of  upkeep  is  rather  high  when  compared  with 
chain  grates. 

Fig.  73  shows  an  application  of  an  American  underfeed  stoker  to  a 
return  tubular  boiler.     This  differs  from  the  Jones  stoker  in  the  method 


FIG.  73.   American  Underfeed  Stoker. 

of  feeding  the  fuel  to  the  retort  and  in  the  employment  of  "  live  " 
grates  instead  of  dead  plates  on  the  sides  of  the  retort.  The  coal  is  fed 
into  the  hopper  and  carried  by  an  endless  screw  conveyor  into  the 
magazine  or  retort.  Forced  draft  is  used  and  the  rate  of  draft  and  the 
speed  of  the  conveyor  are  readily  adjusted  to  suit  the  conditions  of 
load.  Underfeed  stokers  are  very  compact,  occupy  but  little  space  in 
front  of  the  boiler,  and  are  low  in  first  cost.  Careful  manipulation 
is  necessary  to  render  them  smokeless  and  efficient,  since  they  are 
ordinarily  installed  without  fire-tiled  combustion  chambers. 

93.  Down-Draft  Furnaces.  —  Fig.  74  shows  the  application  of  a 
Hawley  down-draft  furnace  to  a  Heine  water-tube  boiler.  In  this 
furnace  there  are  two  separate  grates,  one  above  the  other,  the  upper 


140 


STEAM  POWER  PLANT  ENGINEERING 


one  being  formed  of  paral- 
lel water  tubes  connected 
with  the  water  space  of 
the  boiler  through  the 
steel  headers  or  drums  A 
and  D,  in  such  a  manner 
as  to  insure  a  positive 
circulation.  Fuel  is  sup- 
plied to  the  upper  grate, 
the  lower  one,  formed  of 
common  bars,  being  fed 
by  the  half -consumed  fuel 
falling  from  the  upper 
grate.  Air  for  combustion 
enters  the  upper  fire  door, 
which  is  kept  open,  and 
passes  first  through  the 
bed  of  green  fuel  on  the 
upper  grate  and  then  over 
the  incandescent  fuel  on 
the  lower  grate.  A  strong 
draft  is  required,  due  to 
the  relatively  small  upper- 
grate  area  and  the  corre- 
spondingly high  rate  of 
combustion.  The  down 
draft  is  very  well  adapted 
to  the  burning  of  paper, 
cardboard,  excelsior,  wood , 
and  other  rapidly  burning 
refuse,  as  well  as  the  vari- 
ous kinds  of  coal.  Lump 
coal  gives  better  results 
than  the  smaller  sizes,  as 
the  latter  are  apt  to  fall 
through  the  upper  grate 
before  even  partially  con- 
sumed and  when  such  is 
the  case  efficient  results 
cannot  be  obtained.  If 
carefully  manipulated  this  furnace  with  fire-tiled  tubes  as  illustrated 
in  Fig.  74  gives  high  boiler  efficiency  and  smokeless  combustion, 


SMOKE  PREVENTION,  FURNACES,  STOKERS  141 

but  its  overload  capacity  is  limited.  Without  the  fire  tiling  smokeless  com- 
bustion is  possible  only  at  light  loads.  Hawley  down-draft  furnaces  are  in- 
stalled in  a  large  number  of  tall  office  buildings  in  New  York  and  Chicago 
and  are  giving  excellent  results.  Down-draft  furnaces  are  necessarily 
hand  fired,  since  mechanical  stokers  are  not  easily  adapted  to  them. 

94.  Sprinkling  Stokers.  —  In  this  system  of  stoking  the  fuel  in  finely 
divided  form  is  distributed  by  sprinkling  uniformly  over  the  entire 
area  of  the  grate.     With  the  proper  adjustment  of  air  supply  and  feed 
the   volatile   gases  are   distilled   off   continuously   before  the   grate  is 
covered  by  the  new  .coal  and  without  materially  lowering  the  tempera- 
ture of  the  incandescent  fuel  bed.     Mechanically  the  operation  involves 
considerable  difficulty.     One  of  the  most  successful  American  devices 
for  this  purpose  is  the  "  Little  Giant  "  stoker.     Coal  of  nut  size  or 
smaller  is  hand  fed  into  a  small  hopper  from  which  it  gravitates  on  to  a 
feed  wheel  driven  by  an  engine  or  motor.     A  stream  of  coal  is  dis- 
charged into  a  cast-iron  chute  extending  over  the  front  part  of  the 
grate,  from  which  it  is  blown  into  the  furnace  by  a  steam  jet.     The  fine 
or  powdered  coal  is  burned  in  suspension  and  the  heavier  coal  falls  to 
the  grates.     A  fire-brick  combustion  chamber  is  usually  necessary  for 
smokeless  combustion,  since  the  quantity  of  fuel  burned  may  be  too 
large  to  permit  complete  distillation  of  the  green  fuel  before  the  suc- 
ceeding charges  are  delivered.     A  test  on  an  English  stoker  of  this  type 
(Bennis  stoker), using  compressed  air  for  feeding  the  fuel,  gave  the  unusu- 
ally high  rate  of  combustion  of  72  pounds  of  coal  per  square  foot  of  grate 
surface  per  hour.     The  same  test  credited  the  boiler  and  grate  with  an 
efficiency  of  84.9  per  cent,  which  is   probably  the   highest   recorded 
efficiency  of  any  boiler  and  furnace  using  coal  as  fuel.       (Engineering 
Record,  April  8,  1905,  p.  404.) 

95.  Dutch   Ovens.  —  An    independent    furnace   or    Dutch   oven   in 
front  of  the  boiler  as  illustrated  in  Fig.  75  provides  one  of  the  simplest 
methods  of  securing  a  large  combustion  chamber  for  the  mingling  of 
the  air  and  combustible  gases  before  delivering  them  to  the  boiler  proper. 
Such  a  furnace  produces  very  high  temperatures  when  operating  under 
best  conditions,  and  hence  must  be  lined  with  fire  brick  of  excellent 
quality.     Although  better  than  the  ordinary  setting  the  plain  Dutch 
oven  is  too  limited  in  length  and  capacity  to  prevent  smoke  from  form- 
ing, except  at  very  light  loads.     The  velocity  of  the  gases  is  usually 
too  high  to  permit  either  a  thorough  mixture  or  complete  oxidation 
before  striking  the  boiler  tubes.     Steam  jets  placed  at  the  sides  of  the 
setting  and  blowing  across  the  fire  are  sometimes  effective  in  mixing 
the  air  and  combustible  gases,  but   the  best   results  are  obtained  by 
modifying  the  construction  of  the  furnace  to  the  extent  of  introducing 


142 


STEAM   POWER  PLANT  ENGINEERING 


baffle  walls  which  vary  the  direction  of  flow  and  by  increasing  the 
length  of  the  path  of  the  heated  gases.  The  greater  the  length  of  the 
path  and  the  greater  the  number  of  baffles  the  more  thoroughly  will  the 
air  and  gases  be  mingled,  but  the  intensity  of  draft  will  of  course  be  de- 
creased in  proportion.  A  compromise  must  therefore  be  made  between 
required  draft  and  length  of  path.  The  larger  the  extent  of  fire-tile  sur- 
face the  greater  will  be  the  regenerative  effect,  which  is  of  particular  im- 
portance in  hand  firing  when  the  evolution  of  volatile  gas  is  intermittent, 
but  the  first  investment  and  cost  of  repairs  and  renewals  are  greater. 
There  are  little  reliable  data  available  pertaining  to  the  relation  between 
capacity  of  furnace  and  length  of  path  of  the  heated  gases  for  maximum 


FIG.  75.     Plain  Dutch  Oven. 

efficiency.  A  modified  Dutch  oven  is  illustrated  in  Fig.  71.  The  exten- 
sion front  is  not  necessary  with  some  type  of  boilers,  as  will  be  seen 
from  Figs.  63  and  74,  in  which  a  tile  roof  and  baffles  suitably  arranged 
within  the  setting  proper  simulate  the  Dutch  oven  effect.  (See  "  Cost 
of  Maintenance  of  Dutch  Oven  Furnaces,"  A.  Bement,  Eng.,  U.  S., 
July  1,  1907,  p.  606.) 

96.  Twin-Fire  Furnace.  —  This  arrangement,  illustrated  in  Fig.  76 
in  connection  with  a  hand-fired  return  tubular  boiler,  is  a  double  furnace 
formed  by  longitudinal  arches  extending  between  bridge  wall  and  fire 
door. 

The  furnaces  are  fed  and  manipulated  alternately,  the  object  being 
to  have  one  furnace  in  a  highly  incandescent  state  while  green  fuel  is 
fed  into  the  other.  Air  is  admitted  both  below  and  above  the  grate, 
and  the  volatile  gases  are  supplied  with  the  necessary  oxygen  for  com- 


SMOKE  PREVENTION,  FURNACES,  STOKERS 


143 


3dld    Q33J 


JJO  MO18 


bustion  before  they  come  in  contact  with  the  comparatively  cool  boiler 
surface. 

The  gases  from  both  fur- 
naces first  pass  into  a  cham- 
ber formed  by  a  single  arch 
sprung  across  the  entire  inner 
setting  from  the  side  wall,  a 
short  retarding  arch  being 
placed  between  this  interme- 
diate chamber  and  the  rear 
of  the  setting.  A  special  tile 
of  high-grade  refractory  clay 
is  used,  the  thickness  varying 
from  4  to  6  inches,  depending 
upon  the  size  of  furnace  and 
the  length  of  span.  The  fur- 
nace can  readily  be  substi- 
tuted for  the  ordinary  types 
in  common  use  under  any 
standard  tubular  or  water- 
tube  boiler  and  may  be  in- 
stalled either  under  the  boiler, 
as  indicated  in  the  illustra- 
tion, or  in  an  extension  Dutch 
oven.  This  is  an  excellent 
furnace,  and  when  properly 
manipulated  gives  smokeless 
and  efficient  combustion. 

96a.  Chicago  Settings  for 
Hand-Fired  Return  Tubular 
Boilers.  —  Figs.  67a,  67b  and 
67c  show  the  general  details 
of  settings  for  return  tubular 
boilers  as  recommended  by 
the  Chicago  Department  of 
Smoke  Inspection.  The  set- 
ting illustrated  in  Fig.  67a  is 


ordinarily  installed   where  a 

strong  draft  is  available  and 

that  shown  in  Fig.  67b  or  67c 

where  the  draft  conditions  are  not  favorable.     All  three  settings  require 

careful  manipulation  for  smokeless  combustion  as  is  the  case  with  hand- 


144 


STEAM  POWER  PLANT  ENGINEERING 


1| 


1 


SMOKE  PREVENTION,  FURNACES,  STOKERS 


145 


146 


STEAM  POWER   PLANT   ENGINEERING 


SMOKE  PREVENTION,  FURNACES,  STOKERS  147 

fired  furnaces  in  general.  It  has  been  the  experience  of  the  Department 
that  most  violations  of  the  smoke  ordinance  are  due  primarily  to  insuffi- 
cient draft,  the  required  rate  of  combustion  being  too  high  for  the 
"available  air  supply.  The  requirements  outlined  in  paragraph  95  apply 
equally  well  to  these  settings.  The  following  specifications  refer  to 
Figs.  67a,  67b  and  67c,  the  items  in  the  specifications  corresponding 
to  the  letters  in  the  illustrations. 

A.  Doors  should  be  of  a  type  allowing  the  admission  of  excess  air  over 

the  fire  when  so  desired.  If  panels  are  cut  in  the  fire  doors  for 
this  purpose,  the  aggregate  area  of  the  openings  should  be  not 
less  than  4  square  inches  to  each  square  foot  of  grate  surface. 

B.  Arches  should  be  made  of  wedge  brick  or  "  bull  heads  "  and  not 

laid  in  two  courses  of  4j-inch  brick. 

C.  The  bridge  wall  should  be  made  of  first  grade  fire  brick  above  the 

grate  line  and  with  fire  brick  facing  not  less  than  9  inches 
in  thickness  on  the  combustion  chamber  side.  The  top  row 
should  be  a  row-lock  course.  Provision  should  be  made  in 
the  building  of  the  bridge  wall  for  lateral  expansion. 

D.  The  combustion  chamber  floor  should  be  paved  with  fire  brick 

laid  on  edge. 

E.  Fire  brick  lining  below  the  arch  skew-backs  should  be  not  less 

than  9  inches  in  thickness.  Fire  brick  lining  above  the  arch 
system  and  behind  the  deflection  arch  may  be  4J  inches  in  first 
grade  fire  brick,  with  headers  every  fifth  row. 

F.  Fire  brick  over  firing  door  liners  should  be  arched.     This  rule 

also  applies  to  brick  above  the  clean-out  door  openings. 

G.  Facilities  for  taking  up  arch  thrust  should  be  provided  in  every 

case  by  suitable  metal  re-enforcements  extending  horizontally 
throughout  the  length  of  the  arches.  No  air  space  should  inter- 
vene between  the  metal  re-enforcement  and  the  skew-backs. 

H.  Herringbone  or  Tupper  grates  or  other  similar  types  should  not 
be  selected  where  bituminous  coal  forms  the  major  portion  of 
the  fuel. 

I.  The  back  arch  is  preferably  sprung  from  side  to  side  rather  than 
from  back  wall  to  rear  boiler  tube  sheet.  No  metal  should  be 
exposed  to  direct  heat  of  gases. 

J.  Chimney  heights  of  less  than  75  feet  above  the  grate  line  should 
not  be  permitted,  and  this  height  allowed  only  when  the 
chimney  is  direct  connected  to  the  boiler  uptake.  In  case  of 
a  breeching  and  detached  chimney,  add  to  the  height  of  chim- 
ney computed  by  standard  methods  (never  less  than  75  feet) 
10  feet  for  every  turn  of  the  breeching  and  one  foot  for  each 
foot  in  length  of  the  breeching. 


148  STEAM  POWER  PLANT  ENGINEERING 

K.  For  boilers  48  inches  or  less  in  diameter,  special  provision  for 
the  examination  of  girth  seams  must  sometimes  be  made. 
This  is  because  of  the  fact  that  with  small  boilers  there  is  not 
sufficient  room  between  the  arch  and  shell  for  purposes  of 
inspection. 

UPTAKE 


LONGITUDINAL  SECTION 


SECTIONAL  PLAN 

FIG.  77.     Woolev  Smokeless  Furnace. 


In  the  event  of  arch  failures,  the  boiler  should  be  immediately 
taken  out  of  service.  This  is  to  avoid  failure  of  the  boiler 
shell  due  to  heat  being  applied  upon  a  portion  of  the  heating 
surface  over  which  a  mud  deposit  has  formed. 


SMOKE  PREVENTION,  FURNACES,  STOKERS 


149 


97.  Wooley  Smokeless  Furnace.  —  Fig.  77  shows  a  longitudinal  sec- 
tion and  a  sectional  plan  of  a  Wooley  smokeless  furnace  applied  to  a 
B.  &  W.  boiler.  The  main  features  of  the  furnace  are  a  dividing  wall 
in  the  fire  box  and  a  deflecting  wall  in  the  combustion  chamber.  The 
dividing  wall  permits  of  the  alternate  method  of  firing,  whereby  one 
side  of  the  furnace  is  always  in  an  incandescent  state  while  the  other 
side  is  being  supplied  with  green  fuel.  If  a  mechanical  stoker  is  used 
the  wall  in  the  fire  box  is  omitted.  The  products  of  combustion  are 
intended  to  be  thoroughly  mingled  with  the  requisite  amount  of  air  by 
the  deflecting  walls  before  entering  the  regenerative  or  secondary  com- 
bustion chamber. 


FIG.  78.     Kent's  Wing  Wall  Furnace. 

98.  Kent's  Wing- Wall  Furnace.  —  Fig.  78  shows  the  application  of 
Kent's  wing- wall  furnace  to  a  water-tube  boiler.  The  Dutch  oven 
in  front  of  the  regular  setting  contains  the  grates.  Wing  walls  F  are 
placed  as  shown  two  or  three  feet  to  the  rear  of  the  bridge  wall  D,  and 
fire-brick  piers  H  behind  the  wing  walls. 

In  operation,  fresh  coal  is  spread  alternately  over  each  half  of  the 
grate.  The  dense  smoky  gases  which  rise  from  the  green  portion  of 
the  fire  mingle  in  the  narrow  passage  with  the  highly  heated  air  which 
comes  through  the  other  side  of  the  grate  greatly  in  excess  of  that 
required  to  consume  the  partially  burned  coal  there.  The  piers  H  act 


150 


STEAM  POWER  PLANT  ENGINEERING 


as  regenerative  surfaces,  absorbing  heat  from  the  fire  when  it  is  hottest 
and  giving  it  out  when  it  is  coolest,  that  is,  just  after  firing. 


FIG.  79.    Burke's  Smokeless  Furnace,  Front  Section. 


FIG.  80.     Burke's  Smokeless  Furnace,  Side  Section. 

Comparative  tests  of  boilers  with  standard  setting  and  with  wing-wall 
furnaces  have  shown  a  much  higher  efficiency  with  the  latter  and  with 
practically  smokeless  combustion.  (Iron  Trade  Review,  July  7,  1904, 
p.  76.) 


SMOKE  PREVENTION,  FURNACES,  STOKERS 


151 


99.  Burke's   Smokeless  Furnace.  —  Figs.  79  and  80  show  sections 
through  a  Burke  smokeless  furnace  as  installed  in  a  number  of  tall 
office  buildings  in  Chicago.     It  amounts   virtually  to  a  Dutch  oven 
equipped  with  shaking  grates,  and  embodies  an  extension  self-feeding 
coking  oven  of  cast-iron  section  lined  with  fire  brick  and  protected 
from   overheating   by   air   circulation   through   the  sections.     Natural 
draft  is  used,  the  fire  doors  being  closed;  but  air  is  admitted  above  as 
well  as  below  the  fire.     As  this  stoker  is  manipulated  by  hand,  more  or 
less  attention  is  required  of  the  operator  in  keeping  the  fire  clean. 
Furnaces  of  this  type  at  the  power  plant  of  the  Majestic  Theatre  building, 
Chicago,  111.,  are  giving  excellent  results. 

100.  Admission  of  Air  above  Fire.  —  Smoke  is  often  due  to  insuf- 
ficient air  supply  or  imperfect  mixing,  especially  when  coal  of  a  coking 
or   clinkering   nature  is   used 

which  tends  to  seal  up  the  air 

spaces  in  the  grate.     In  these 

cases  the  admission  of  air  above 

the  grate  through  openings  in 

the  bridge  wall  or  passages  in 

the  side  walls  frequently  gives 

satisfactory     results.       When 

natural  draft  is  not  sufficient, 

as  is  usually  the  case  under 

heavy  load,  steam  jets  or  forced   draft   may  be  employed.     For  a 

description  of  such  devices  see  paragraphs  148  to  150. 

101.  Cost  of  Stokers. — The  following  is  the  approximate  cost  of  stokers 
suitable  for  a  Babcock  &  Wilcox  boiler  of  350  horse  power  rated  capacity 
with  45  square  feet  of  grate  surface ;  height  of  chimney  above  grate,  1 75  feet ; 
coal  burned,  Illinois  screenings.     The  cost  of  installation  is  not  included. 

1.  Chain  grate  and  appurtenances $1,500.00 

2.  Jones  underfeed  stoker 1,400.00 

3.  Hawley  down-draft  furnace 1,350.00 

4.  Burke  smokeless  furnace 1,000.00 

5.  Roney  stoker 1,300.00 

6.  Murphy  furnace  and  stoker 1,350.00 

7.  Wilkinson  stoker 1,200.00 

Parsons'  Smokeless  Furnace.     See  par.  149. 

Heinrich  Smokeless  Furnace.     See  par.  150. 
Steam  jets.     See  par.  148. 
Hamler-Eddy  Smoke  Recorder.     See  par.  41  la. 
Ringlemann  Smoke  Chart.     See  p.  765. 

Smoke  Prevention:  Bulletin  No.  15,  Univ.  of  111.,  Vol.  431 ;  Bulletin  No.  334,  U.  S. 
Geological  Survey;  Boiler  Maker,  May,  1909,  Oct.,  1909;  Cassier's  Mag.,  Feb.,  1907; 
Minn.  Engr.,  Jan.,  1910. 

Mechanical  Stokers:  Engr.  U.  S.,  Jan.  1, 1907,  p.  83,  Aug.  15, 1906,  p.  540,  July  2, 1906, 
p.  437;Cassier's  Mag.,  Sept.,  1906,  p.  469;  Power, Mar.,  1906,  p.  189,  Aug.,  1905,  p.  487. 


AIR  FROM  ASM-PAN 


FIG.  81.     Split  Bridge  Wall. 


CHAPTER  V. 

SUPERHEATED  STEAM;  SUPERHEATERS. 

102.  General.  —  The  steam  engine  fails  to  realize  the  efficiency  of 
the  ideal  engine  chiefly  on  account  of  cylinder  condensation.  The  loss 
in  heat  due  to  this  cause  is  seldom  less  than  10  per  cent  of  the  total 
supplied,  and  often  as  great  as  40  per  cent. 

If  the  steam  is  superheated  before  being  admitted  to  the  cylinder, 
condensation  may  be  reduced  or  prevented  entirely,  as  was  recognized 
as  early  as  sixty  years  ago,  but  the  mechanical  difficulties  encountered 
prevented  the  practice  until  within  the  past  few  years. 

The  principal  advantages  of  superheated  steam  in  connection  with 
steam-engine  work  are: 

1.  At  high  temperatures  it  behaves  like  a  gas  and  is  therefore  in  a 
far  more  stable  condition  than  in  the  saturated  form.  Considerable 
heat  may  be  abstracted  without  producing  liquefaction,  whereas  the 
slightest  absorption  of  heat  from  saturated  steam  results  in  condensa- 
tion. If  superheat  is  high  enough  to  supply  not  only  the  heat  absorbed 
by  the  cylinder  walls  but  also  the  heat  equivalent  of  the  work  done 
during  expansion,  then  the  steam  will  be  dry  and  saturated  at  release. 
This  is  the  condition  of  maximum  efficiency  in  a  single  cylinder. 
(Ripper,  "  Steam  Engine  Theory,"  p.  155.)  Greater  superheat  than  this 
will  result  in  a  loss  of  energy  unless  the  steam  is  exhausted  into 
another  cylinder.  To  obtain  dry  steam  at  release  the  steam  at  cut  off 
must  be  superheated  100  to  300  degrees  F.  above  saturation  tempera- 
ture, depending  upon  the  initial  condition  of  the  steam  and  the  number 
of  expansions,  a  higher  degree  of  superheat  being  required  for  earlier 
cut  off.  A  superheat  of  200  to  275  degrees  F.  at  admission  is  necessary 
to  insure  dry  steam  at  release  in  the  average  single-cylinder  engine 
cutting  off  at  one-fourth  stroke,  boiler  pressure  100  pounds  gauge.  In 
most  cases  superheat  is  only  carried  so  far  as  to  reduce  initial  conden- 
sation, the  steam  becoming  saturated  at  cut  off,  thus  permitting  efficient 
lubrication.  There  will  be  a  reduction  of  approximately  1  per  cent  in 
cylinder  condensation  for  every  7.5  to  10  degrees  of  superheat.  In 
compound  and  triple-expansion  engines  the  steam  is  ordinarily  super- 
heated between  each  stage  as  well  as  before  admission  to  the  high- 
pressure  cylinder. 

152 


SUPERHEATED  STEAM;  SUPERHEATERS        153 

2.  .A  moderate  amount  of  superheat   produces  a  large  increase  in 
volume,  the  pressure  remaining  constant,  and  diminishes  the  weight  of 
steam  per  stroke  for  a  given  amount  of  work.     For  example,  the  volume 
of  1  pound  of  saturated  steam  at  150  pounds  pressure  (gauge)  is  2.75 
cubic  feet,  and  its  temperature  is  365.8  degrees  F.    The  total  heat  of  one 
pound  of  this  steam  above  the  freezing  point  is  1193.5  B.T.U.     By 
adding  110  B.T.U.  in  the  form  of  superheat  its  temperature  will  be 
increased  to  565.8  degrees  F.  (superheated  200  degrees  F.)  and  its  volume 
to  approximately  3.5  cubic  feet  (specific  heat  taken  as  0.55).*     Thus 
an  increase  of  9.2  per  cent  in  the  heat  effects  an  increase  of  22  per  cent 
in  the  volume,  which  means  a  corresponding  reduction  in  the  steam 
admitted  to  the  engine  per  stroke.     These  figures  are  purely  theoretical, 
as  no  allowances  have  been  made  for  condensation  of  the  saturated 
steam  or  for  reduction  in  temperature  of  the  superheated  steam. 

3.  Superheated    steam    has   a   much   lower    thermal    conductivity 
than  saturated  steam,  and  therefore,  less  heat  is  absorbed  per  unit  of 
time  by  the  cylinder  walls. 

General  Discussion  of  Superheated  Steam:  Engr.,  Lond.,  Dec.  31,  1909;  Eng., 
Lond.,  Sept.  13,  1901,  Sept.  4,  1903,  p.  237;  Eng.,  U.S.,  Dec.  15,  1902,  p.  821, 
Oct.  15,  1906,  p.  687;  Engr.  Mag.,  Feb.  1903,  p.  778,  Sept.,  1903,  p.  897,  Feb., 
1904,  p.  757,  June,  1904,  p.  436,  March,  1905,  p.  943,  Nov.,  1905,  p.  271,  May, 
1906,  p.  269;  Eng.  Rec.,  July  8,  1905,  p.  28;  June  30,  1906,  p.  783,  July  28,  1906, 
p.  86;  Power,  Aug.,  1904,  p.  463,  Sept.,  1904,  p.  558,  Oct.,  1904,  p.  762,  Jan.,  1905, 
p.  23,  Feb.,  18,  1908,  Serial;  Eng.,  Lond.,  Jan.  8,  1904,  p.  42;  West.  Elec.,  Nov. 
14,  1903,  p.  369;  Proc.  A.S.M.E.,  May  14,  1908. 

103.  Economy  of  Superheat.  —  Many  comparative  tests  of  engines 
using  saturated  and  superheated  steam  under  varying  conditions  of 
pressure  and  temperature  have  been  made  during  the  past  few  years, 
showing  in  most  cases  a  gain  in  favor  of  superheat  due  to  the  reduction 
in  steam  consumption,  but  in  some  cases  the  extra  investment  and  cost 
of  maintenance  neutralize  this  gain,  resulting  in  an  actual  loss  when 
measured  in  dollars  and  cents  per  horse-power  hour. 

As  far  as  steam  consumption  per  horse-power  hour  is  concerned, 
superheating  usually  increases  the  economy  five  to  fifteen  per  cent 

*  The  most  satisfactory  equation  for  determining  the  specific  volume  of  super- 
heated steam  is  that  given  by  Knoblauch,  Linde,  and  Klebe  (Peabody,  "  Steam  and 
Entropy  Tables,"  p.  22): 


pv  =  0.5962  T  -  p(l  +  0.0014  p)  _  ao833 


p  —  pressure,  pounds  per  square  inch  absolute. 
T  =  absolute  temperature  of  the  steam,  degrees  F. 
v  =  specific  volume  of  superheated  steam,  cubic  feet. 


154  STEAM  POWER  PLANT  ENGINEERING 

and  in  some  instances  as  much  as  forty,  the  latter  figure  referring  to 
the  more  wasteful  types  of  engines.  A  fair  estimate  of  the  average 
reduction  in  steam  consumption  per  horse-power  hour  with  moderate 
superheating,  that  is,  100  to  125  degrees  F.,  based  on  continuous  opera- 
tion of  existing  plants,  is : 

Per  Cent. 

1.  Slow  running,  full  stroke,  or  throttling  engines,  including  direct 

acting  pumps 40 

2.  Simple  engines,  non-condensing,  with  medium  piston  speed,  includ- 

ing compound  direct  acting  pumps 20 

3.  Compound  condensing  Corliss  engines 10 

4.  Triple-expansion  engines 6 

A  prominent  European  builder  of  engines  guarantees  steam  con- 
sumption with  highly  superheated  steam  as  follows: 

Pounds  per  I.H.P.  hour. 

Single-cylinder  condensing  engines 13.5 

Single-cylinder  non-condensing  engines, 15.5 

Compound  condensing  engines 10 

Triple-expansion  condensing  engines 8.75 

In  comparing  the  performances  of  engines  using  saturated  and 
superheated  steam  it  is  advisable  to  base  all  results  on  the  heat  con- 
sumed per  horse  power  rather  than  on  the  steam  consumption,  since 
the  latter  is  apt  to  give  a  false  idea  of  the  relative  economies.  The 
real  measure  of  economy  is  the  cost  of  producing  power,  taking  into 
consideration  all  charges,  fixed  and  operating,  and  the  next  best  is  the 
coal  consumption  per  I.H.P.  hour,  but  as  a  means  of  comparing  the 
engines  only,  the  heat  consumption  per  horse  power  per  hour  or  per 
minute  is  very  satisfactory. 

See  paragraph  181  for  the  influence  of  superheat  on  the  economy  of 
reciprocating  engines  and  paragraph  193  for  the  influence  on  steam 
turbines. 

Economy  of  Superheat :  Eng.  Mag.,  Dec.,  1904,  p.  757,  April,  1905,  Sept.,  1903, 
p.  108;  Trans.  A.S.M.E.,  22-899;  Engr.  Rec.,  July  8,  1905,  p.  28;  Power,  Sept., 
1904,  p.  558,  Oct.,  1904,  p.  598,  Jan.,  1905,  p.  23;  Cassier's,  Nov.,  1903,  p.  18. 

104.  Limit  of  Superheat.  —  In  this  country  steam  temperatures 
exceeding  500  degrees  F.  are  seldom  employed,  while  in  Europe  few 
if  any  plants  are  installed  without  superheaters,  and  600  degrees  F. 
is  a  common  temperature. 

Experience  has  shown  that  with  engines  of  ordinary  design,  slide- 
valves  and  Corliss,  the  temperature  at  the  throttle  should  not  exceed 
500  degrees  F.  This  corresponds  to  a  superheat  of  160  degrees  F.  with 


SUPERHEATED  STEAM;  SUPERHEATERS        155 

steam  at  100  pounds  gauge  pressure,  and  130  degrees  F.  at  150  pounds. 
This  degree  of  superheat  insures  practically  dry  steam  at  cut  off  in  the 
better  grade  of  engines.  Just  how  far  superheating  can  be  carried  with 
a  given  engine  of  ordinary  construction  can  be  determined  by  experiment 
only,  but  a  temperature  of  500  degrees  F.  is  probably  an  outside  figure 
and  450  degrees  F.  a  good  average.  Higher  temperatures  are  apt  to 
interfere  with  lubrication  and  sometimes  cause  warping  of  the  valves. 
With  temperatures  below  450  degrees  F.  no  difficulties  are  ordinarily 
met  with.  Metallic  packing  has  been  found  to  give  the  best  results  for 
both  piston  rods  and  valve  stem. 

It  is  generally  assumed  that  a  greater  quantity  of  oil  is  required 
for  lubricating  valves  and  cylinders  in  connection  with  superheated 
steam,  but  experience  seems  to  show  that  such  is  not  the  case.  (Proc. 
A.S.M.E.,  May  14,  1908.)  Forced-feed  lubricators  are  the  most  satis- 
factory for  superheated  steam  engines,  since  they  insure  a  positive 
and  copious  flow  of  oil  directly  to  the  valves  or  other  parts  requiring  it.* 

With  highly  superheated  steam  involving  temperatures  of  600 
degrees  F.  or  more  the  poppet-valve  type  of  engine  is  ordinarily 
employed,  though  balanced  piston  valves  are  not  uncommon.  The 
poppet  valve  is  not  distorted  by  heat  and  requires  no  lubrication.  In 
Europe  these  engines  have  been  brought  to  a  high  state  of  efficiency, 
but  have  not  been  generally  adopted  in  this  country  owing,  no  doubt, 
to  the  higher  cost. 

105.  Specific  Heat  of  Superheated  Steam,  f  —  The  total  heat  of  super- 
heated steam  is  given  as  „  ,  .  n  f  /i  7\ 

tl  —  A  +  Opt,  \i  I  ) 

in  which 

A  =  B.T.U.  in  one  pound  of  saturated  steam  above  32  degrees  F. 

Cp=  mean  specific  heat  of  the  superheated  steam  at  constant  pressure. 
t  =  degree  of  superheat,  degrees  F. 

Regnault  determined  the  mean  specific  heat  at  atmospheric  pressure 
to  be  0.48  between  127  degrees  and  226  degrees  C.  of  superheating,  and 
until  recently  this  has  been  assumed  to  apply  to  all  pressures  and  tem- 
peratures. As  early  as  1876  Hirn  concluded  from  experiments  made 
with  a  throttling  calorimeter  that  the  specific  heat  of  saturated  steam 
increased  with  the  pressures  and  decreased  at  any  given  pressure  if  the 
steam  became  superheated.  Since  then  numerous  investigators  have 
promulgated  theories  pertaining  to  this  subject  which  have  been  far 
from  harmonious  and  none  has  been  universally  accepted.  Some 
experiments  appear  to  show  that  specific  heat  is  independent  of  pressure 

*  Effect  of  Superheated  Steam  on  Cylinder  Oils.  Mech.  Engr.,  Lond.,  July  31, 
1908,  p.  115. 

t  See  paragraph  113a. 


156 


STEAM   POWER  PLANT  ENGINEERING 


and  degree  of  superheat,  while  others  indicate  an  increasing  value  as 
the  pressure  and  degree  of  heat  increase.  Still  others  corroborate 
Hirn's  theory. 


0.46 


200  300  400  500  600  700 

Temperature  -Deg.  Eab 

FIG.  82.   Specific  Heat  of  Superheated  Steam,    Knoblauch  and  Jakob. 


.8 


40 


00 


140          160  180         300 


80  100  120 

Pressure-Lb.  Abs. 
FIG.  83.    Specific  Heat  of  Superheated  Steam,  A.  R.  Dodge. 

The  maximum  figure  ranges  as  high  as  0.8  and  the  minimum  0.48  for 
a  given  pressure  and  degree  of  superheat. 

The  curves  in  Fig.  82  are  based  upon  the  experiments  of  Knoblauch 
and  Jakob  ("  Mitteilungen  iiber  Forschungsarbeiten,"  etc.,  Heft  36, 
p.  109,  and  Stevens'  Indicator,  October,  1905);  those  in  Fig.  83  upon  the 


SUPERHEATED  STEAM;  SUPERHEATERS 


157 


experiments  of  A.  R.  Dodge  (Trans.  A.S.M.E.,  1907);  those  in  Fig.  84 
are  plotted  from  tests  of  Burgoon,  Carpenter,  and  Thomas  (Trans. 
A.S.M.E,  1907);  and  those  in  Fig.  85  are  based  upon  the  investigation 
of  Professor  Thomas  (Trans.  A.S.M.E.,  December,  1907).  These  curves 


470 


Specific  Heat  of  Steam 
Curves  in  Black  :-C.E.Burgoon 
"       »  Light  Lines:- Callendar 
Marked    CO 


120          160  200  240          280 

Degrees,  F.  of  Superheat 
FIG.  84.     Specific  Heat  of  Superheated  Steam,  C.  E.  Burgoon. 

differ  both  in  theory  and  in  value  of  cp,  but  until  further  experiments 
prove  otherwise  the  values  in  Fig.  82  may  be  accepted  as  sufficiently 
accurate  for  all  practical  purposes.  The  values  given  by  Knoblauch 
and  Jakob  have  been  accepted  by  authorities  as  the  most  reliable. 
Table  17a  is  based  upon  their  results.  Table  17b  has  been  calculated 
by  means  of  Linde's  equation.  (See  footnote,  page  153.) 

TABLE  17.  . 

VALUE   OF  cp  AT  ATMOSPHERIC  PRESSURE  BY    VARIOUS  AUTHORITIES. 
Superheated  Steam  Cooled  by  Water- Jacketed  Calorimeter. 


Author. 

Publication  and  Date 

Temp. 

cp  at 

Variation 

of  cp  with 

Deg.  F. 

Pres. 

Increasing 
Pressure. 

Increasing 
Temp. 

Regnault  

Carpenter  (Jones) 
Dodge  

Ann.  de  Chimie 
et  de  Physique, 
Tome  23. 

Sibley  Journal, 
5-1904. 
Trans  A  S  M  E 

Varied 
Varied 

0.4805 

0.4844 
0  48 

None 
Increases 

None 
None 

158 


STEAM   POWER  PLANT  ENGINEERING 


TABLE    17  —  Continued. 
Throttling  Calorimeter.     Saturated  Steam  Expanded  to  Lower  Pressure. 


Author. 

Publication  and  Date. 

Temp. 
Deg.  F. 

cp  at 
Atmos. 
Pres. 

Variation  of  cp  with 

Increasing 
Pressure. 

Increasing 
Temp. 

Grindley  

Phil.  Trans.,  Vol. 
194. 

239 

0.4317 

Increases 

Increases 
Increases 

Increases 
Increases 
Increases 
Increases 

Increases 

Decreases 
Increases 

Increases 
None 
None 
None 

Him  

Griessmann  

Zeit.  V.  D.  Ing.,  52, 
1903. 
Proc.  Royal  See. 
A-509,  1905. 
Sibley  Journal, 
May,  1904. 
Sibley  Journal, 
May,  1904. 
Sibley  Journal, 
May,  1904. 

269 
Varied 
Varied 
Varied 
Varied 

0.506 
0.43 
0.463 
0.4825 
0.48 

Peake 

Carpenter  (Stew- 
art and  Marble).. 
Carpenter  (Hoxie 
and  Wood). 
Carpenter 
(Sickles). 

Superheating  Steam  Electrically. 

Peake       

Proc.  Royal  Society 
A509,  1903. 

Varied 

Varied 
212 
402 

'   212 
700 

0.46 

0.48 
0.49 
0.487 

0.445 
0.49 

None 
Increases 

None 
None 

Carpenter  (Berry) 
Carpenter 
(Thomas). 

Lorenz  
Knoblauch  and 
Jakob. 

Trans.  A.S.M.E., 
Eng.   Mag.,  March 
1907. 
Z.V.D.I.,No.  20.... 
Engineering,  L, 
Feb.  22,  1907; 

Increases 
Increases 

Decreases 

Decreases 
Decreases 
then 
increases 

Decreases 

From  Combustion  of  Explosive  Gases. 


Mallard  and 

Zeit.    V.    D.    Ing., 

212 

0.46 

None 

Increases 

LeChatelier. 

Tome  48. 

Sarran  and  Vieille 

.  Do  

212 

0.464 

None 

Increases 

.  .     Do  

212 

0.463 

None 

Increases 

From  Calculation. 


Reeve     

Wor.  Poly.  Journal 

215 

0.39 

Increases 

Increases 

Him     

265 

0.4895 

Increases 

Decreases 

Grav 

London  Engineer  .  . 

236 

0.38 

*     Zeuner               .... 

0.568 

Weyrauch     

Zeit.  V.  D.  Ing., 

212 

0.468 

Decreases 

Increases 

Perry  

Tome  48. 
Steam  Engine  

212 

0.36 

Increases 

None 

Roentgen          .... 

Thermodynamics  .  . 

Varied 

0.4805 

None 

None 

\Vagner        

Rose  Technic,  1905 

284.4 

0.513 

Increases 

Knoblauch,  Linde, 

an/1     T?"1oV-»o 

Publication  by  au- 

tlior^     Horlm   1QO^ 

212 

QKft 

0.493 
0  479 

Increases 

Decreases 

SUPERHEATED  STEAM;  SUPERHEATERS 


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162 


STEAM   POWER  PLANT  ENGINEERING 


Just  how  much  these  different  values  of  the  specific  heat  affect  the 
calculated  performance  of  an  engine  may  be  illustrated  by  the  following 
case: 

An  engine  uses  8.5  pounds  of  steam  per  I.H.P.  hour,  initial  pressure 
120  pounds  absolute;  back  pressure  0.94  pound  absolute;  superheat 
at  throttle  300  degrees  F.  Assume  that  the  "  heat  of  the  liquid  " 
in  the  exhaust  steam  is  returned  to  the  boiler  and  that  the  specific 
heat  is  0.48  in  one  case  and  0.8  in  the  other,  these  being  the  extreme 
values  given  by  different  experimenters  for  the  given  conditions. 


Case  I. 

Case  II. 

Specific  heat  

0  48 

o 

8 

B.T.U.  in  1  pound  of  saturated  steam  above 
ideal  hot-  well  temperature  

1118  0 

1118  0 

B.T.U.  in  superheat  

144 

240 

B.T.U.  in  1  pound  of  superheated  steam  .  . 
B.T  U   per  I.H  P   per  minute 

1262.0 
162  1 

1358.0 
192 

4 

On  the  basis  of  the  specific  heat  of  0.48  the  heat  consumption,  162.1 
B.T.U.  per  I.H.P.  per  minute,  is  exceptional  and  corresponds  to  an 
equivalent  saturated  steam  consumption  of  9.6  pounds  per  I.H.P. 
hour,  whereas  the  performance  of  192.4  B.T.U.  per  I.H.P.  per  minute 
based  upon  the  specific  heat  of  0.8  has  been  excelled  by  a  number  of 
actual  engines  using  saturated  steam. 

106.  Superheaters.  —  The  installation  of  a  superheater  is  equivalent 
to  an  increase  in  boiler  capacity.  The  superheater  may  be  independ- 
ently fired  or  can  be  arranged  in  connection  with  the  boiler  proper. 
The  heating  surface  is  usually  of  wrought  iron,  mild  steel,  cast  iron,  or 
cast  steel.  Engineers  are  not  agreed  as  to  which  arrangement  or  which 
material  gives  the  most  economical  returns.  The  recent  symposium 
on  superheated  steam  conducted  by  the  research  committee  of  the 
American  Society  of  Mechanical  Engineers  (Trans.  A.S.M.E.,  1907) 
clearly  indicated  this  lack  of  agreement.  The  requirements  for  a  suc- 
cessful superheater  are: 

1.  Security  in  operation,  or  minimum  danger  of  overheating. 

2.  Economical  use  of  heat  applied. 

3.  No  exposure  of  joints  to  the  fire. 

4.  Provision  for  free  expansion. 

5.  Disposition  such   that   it  may  be   cut  out   or  repaired  without 
interfering  with  the  operation  of  the  plant. 

6.  Ease  of  application  to  existing  plants. 


SUPERHEATED  STEAM;  SUPERHEATERS 


163 


Nearly  all  superheaters  depend  upon  carrying  steam  at  a  high  velocity 
through  small  tubes  in  the  form  of  return  bends  or  coils  and  arranged 
to  be  heated  by  the  hot  gases  in  the  boiler  furnace  or  from  some  other 
source. 

The  independently  fired  superheater  has  the  following  advantages: 

1.  The  degree  of  superheat  may  be  varied  independently  of  the  per- 
formance of  the  boiler. 

2.  It  can  be  placed  at  any  desirable  point. 

3.  Repairs  are  readily  made  without  shutting  down  boilers. 

Some  of  the  disadvantages  are: 

1.  It  requires  separate  firing  and  extra  attention. 

2.  Saturated  steam  can  only  be  furnished  in  case  of  a  breakdown  to 
the  superheater. 

3.  Extra  piping  is  required. 

4.  Extra  space  required. 

Standard  practice  in  this  country  advocates  that  the  superheater  be 
contained  within  the  boiler  setting.  Of  two  hundred  recent  installa- 
tions, one  hundred  and  eighty,  or  ninety  per  cent,  were  of  this  type. 


FIG.  86.     Babcock  and  Wilcox  Superheater. 

107.  Babcock  &  Wilcox  Superheater.  —  Fig.  86  shows  the  applica- 
tion of  superheating  coils  to  a  Babcock  &  Wilcox  boiler,  illustrating 
the  indirectly  fired  type.  The  wrought-iron  tubes  are  bent  into  U  shape, 
the  ends  being  connected  into  manifolds,  the  upper  one  receiving  the 
saturated  steam  from  the  boiler  and  the  lower  one  the  superheated 


164 


STEAM  POWER  PLANT  ENGINEERING 


steam  after  it  has  traversed  the  superheater  tubes.  A  small  pipe  con- 
nects the  lower  manifold  with  the  water  space  of  the  boiler  by  means  of 
which  the  superheater  may  be  cut  out  if  desired,  or  flooded  when  starting 
up.  Any  steam  formed  in  the  superheater  tubes  is  returned  into  the 
boiler  drum  through  the  collecting  pipe,  which,  when  the  superheater 
is  at  work,  conveys  saturated  steam  into  the  upper  manifold.  When 
steam  pressure  has  been  attained  the  superheater  is  thrown  into  action 
by  draining  the  water  away  from  the  manifolds  and  opening  the  super- 
heater stop  valve.  The  tubes  are  free  at  one  end  and  the  manifolds  are 
not  rigidly  connected  with  each  other,  thus  avoiding  expansion  strains. 
With  the  proportion  of  superheating  surface  to  boiler  surface  ordinarily 
adopted  the  steam  is  superheated  from  100  to  150  degrees  F, 


STEAM  PIPE 


SAFETY  VALVC 


FIG.  87.    Stirling  Superheater. 

108.  Stirling  Superheater.  —  This  superheater  consists  of  two  drums, 
Fig.  88,  connected  by  seamless  drawn  tubes  two  inches  in  diameter.  It 
may  take  the  place  of  the  middle  bank  of  tubes  in  the  Stirling  boiler  as 
shown  in  Fig.  87,  or  be  installed  in  the  final  pass  of  the  gases  in  the  back 


SUPERHEATED  STEAM;  SUPERHEATERS 


165 


of  the  boiler.  The  drums  around  the  tubes  are  protected  from  intense 
heat  by  asbestos  cement.  A  pipe  connecting  the  front  drum  of  the 
boiler  with  the  lower  drum  of  the  superheater  permits  the  coils  to  be 
flooded  in  starting  up  or  when  the  superheater  is  not  needed.  In  this 


FIG.  88.     Arrangement  of  Tubes ;  Stirling  Superheater. 


case  the  superheater  acts  as  additional  boiler-heating  surface.  The 
upper  drum  is  divided  into  three  and  the  lower  into  two  compartments. 
The  tubes  are  arranged  with  alternately  wide  and  narrow  spacing,  so 
that  any  tube  may  be  removed  without  disturbing  the  rest.  The  flow 
of  steam  is  indicated  by  arrows. 

109.  Foster  Superheater.  —  Fig.  89  shows  the  application  of  a 
Foster  superheater  to  a  Babcock  &  Wilcox  boiler.  This  device  consists 
of  cast-iron  headers  joined  by  a  bank  of  straight  parallel  seamless 
drawn-steel  tubes,  each  tube  being  encased  in  a  series  of  annular  flanges 
placed  close  to  each  other  and  forming  an  external  cast-iron  covering 
of  large  surface.  The  tubes  are  double,  the  inner  tube  serving  to  form 
a  thin  annular  space  through  which  the  steam  passes  as  indicated. 
Caps  are  provided  at  the  end  of  each  element  for  inspection  and  cleaning 
purposes.  Foster  superheaters  are  more  costly  than  plain-tube  super- 
heaters, but  are  longer  lived  and  offer  a  much  larger  heating  surface  in 
proportion  to  the  space  occupied. 

Fig.  91  shows  a  Foster  superheater  arranged  for  independent  firing. 


166 


STEAM   POWER  PLANT  ENGINEERING 


The  "  Schwoerer  "  superheater,  which  is  somewhat  similar  in  external 
appearance  to  the  Foster,  differs  from  it  considerably  in  detail,  the 
heating  surface  being  made  up  of  suitable  lengths  of  cast-iron  pipe 
ribbed  outside  circumferentially  and  inside  longitudinally.  The  ends 
of  the  pipes  are  flanged  and  connected  by  cast-iron  U-bends.  The 
intention  is  to  provide  ample  heating  surface  internally  and  externally, 
with  a  compact  apparatus. 


FIG.  89.    Foster  Superheater  in  Babcock  &  Wilcox  Boiler. 

110.  Independently  Fired  Superheaters.  —  The  Schmidt  superheater, 
Fig.  90,  consists  of  two  nests  of  coils,  A  and  D,  of  equal  size  and  dimen- 
sions, connected  to  cast-iron  headers  0  and  7.  Saturated  steam  enters 
the  first  nest  of  coils  through  C  and  passes  into  header  0.  From  0 
the  steam,  which  is  now  dried  and  partly  superheated,  flows  through 
the  cast-iron  pipe  E  to  header  /,  and  thence  through  the  second  nest 
of  coils  into  header  adjoining  0,  and  through  pipe  R  to  the  engine. 
In  chamber  D  the  steam  and  gases  flow  on  the  counter-current  and  in 
chamber  A  on  the  concurrent  principle.  This  combination  permits  of 
a  low  flue  temperature  and  high  steam  temperature  without  subjecting 
the  tubes  to  an  excess  of  heat  as  would  be  the  case  if  the  steam  left  the 
coils  A  at  header  7,  where  the  furnace  gases  are  the  hottest.  A  steam 


SUPERHEATED  STEAM;  SUPERHEATERS 


167 


1 


I 


168 


STEAM   POWER   PLANT   ENGINEERING 


1 


SUPERHEATED  STEAM;  SUPERHEATERS 


169 


temperature  of  750  degrees  F.  and  a  flue  temperature  of  450  degrees  F. 
are  easily  maintained  with  this  apparatus.  A  mercury  pyrometer  T 
is  fitted  where  the  superheated  steam  enters  the  discharge  pipe  R. 
A  thermometer  cup  L  permits  of  checking  the  pyrometer  by  means  of 
a  nitrogen-filled  thermometer.  Each  coil  can  be  taken  out  separately 
and  a  new  one  put  in  without  removing  the  others  or  dismantling  the 
plant.  Water  produced  by  condensation  while  the  superheater  is 
inoperative  collects  in  the  bottom  header  N  and  escapes  through  a 
drain  cock.  If  the  steam  supply  should  be  suddenly  shut  off,  the  air 
door  P  is  opened  automatically  by  weight  K.  As  soon  as  steam 
begins  to  flow  it  raises  the  weight  through  the  opening  of  valve  C  and 
the  door  closes.  The  Schmidt  superheater  when  arranged  in  the  flue 
has  practically  the  same  construction  as  the  independently  fired. 


ECONOMIZER 


FIG.  92.    Schmidt  System  of  Combined  Superheater,  Feed  Water  Heater  and  Economizer. 

Fig.  92  shows  a  combination  of  Schmidt  superheater,  economizer, 
and  feed-water  heater  which  finds  much  favor  with  engineers  on  the 
Continent. 


170  STEAM   POWER  PLANT  ENGINEERING 

111.  Materials  used  in  Construction  of  Superheaters.  —  Most  super- 
heaters are  constructed  either  of  wrought  iron,  mild  steel,  cast  iron,  or 
cast  steel,  the  latter  material  having  the  advantage  of  not  being  dam- 
aged by  any  temperature  to  which  it  is  likely  to  be  subjected,  which 
does  away  with  the  necessity  of  damper  mechanisms  and  simplifies  the 
installation.     On  the  other  hand,  cast-metal  superheaters  are  usually 
ribbed  after  the  fashion  of  an  air-cooled  gas  engine,  and  are,  therefore, 
very  heavy  and  thick  walled,  necessitating  a  higher  temperature  for 
the  same  useful  effect  than  in  the  case  of  the  wrought-iron  construction, 
but  have  the  advantage  of  minimizing  fluctuation  of  steam  temperature 
which  would  otherwise  be  caused  by  a  wide  variation  in  temperature  of 
furnace.     One  of  the  most  successful  cast-metal  heaters  is  of  European 
design  and  is  constructed  of    a  special  alloy  known  as  "  Schwoerer  " 
iron.     Table  18  gives  the  yearly  cost  of  repairs  to  piping  and  necessary 
brickwork  for  a  number  of  installations  equipped  with  cast-metal  super- 
heaters of  the  "  Schwoerer  "  type. 

Wrought  iron  and  mild  steel  offer  the  advantage  of  lightness,  ease  of 
construction,  and  low  first  cost,  but  cannot  be  exposed  to  very  high 
temperatures  without  injury,  and  consequently  provision  must  be  made 
for  diverting  the  direction  of  the  heated  gases  or  for  flooding  the  coils 
while  the  boiler  is  being  warmed  before  steam  is  generated. 

Neither  cast  iron  nor  steel  loses  in  tensile  strength  when  subjected  for 
a  very  short  time  to  the  temperature  of  superheated  steam,  but,  on  the 
contrary,  may  be  stronger.  Tests  made  by  Professor  Lanza  (" Applied 
Mechanics,"  p.  489)  showed  that  the  tensile  strength  of  steel  dimin- 
ished from  0  degrees  F.  to  about  300  degrees  F.  and  then  increased, 
reaching  a  maximum  between  500  and  650  degrees  F.  Cast  iron  and 
steel  maintained  their  strength,  with  a  tendency  to  increase,  up  to  900 
degrees  F.  beyond  which  the  strength  is  diminished. 

Ordinary  cast-iron  valves  and  fittings  have  shown  permanent  increase 
in  dimensions  under  high  superheat  and  in  numerous  instances  have  failed 
altogether,  but  sufficient  data  are  not  available  to  prove  conclusively  the 
unreliability  of  cast  iron  if  the  iron  mixture  is  properly  compounded 
and  the  necessary  provision  is  made  for  expansion  and  contraction.* 

112.  Extent  of  Superheating  Surface.  —  The  required  extent  of  super- 
heating surface  for  any  proposed  installation  depends  upon   (1)   the 
degree  of  superheat  to  be  maintained;  (2)  the  Velocity  of  the  steam 
through  the  superheater;  (3)  the  character  of  the  superheater;  (4)  the 
weight  of  the  steam  to  be  superheated;  (5)  the  moisture  in  the  wet  steam ; 
(6)  the  temperature  of  the  gases  entering  and  leaving  the  superheater; 
and  (7)  the  conductivity  of  the  material. 

*  See  Symposium  on  the  "Effect  of  Superheated  Steam  on  Cast  Iron  and  Steel," 
Jour.  Am.  Soc.  Mech.  Engrs.,  Dec.,  1909. 


SUPERHEATED  STEAM;  SUPERHEATERS 


171 


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172  STEAM   POWER  PLANT  ENGINEERING 

Since  the  heat  absorbed  by  the  steam  in  the  superheater  is  equal  to 
that  given  up  by  the  products  of  combustion,  neglecting  radiation,  this 
relationship  may  be  expressed 

SUd  =  We  ft  -  Z2),  (23) 

in  which 

S  =  square  feet  of  superheating  surface  per  boiler  horse  power. 
U  =  coefficient  of   heat  transmission,  B.T.U.  per  square  foot  per 

hour  per  degree  difference  in  temperature. 
d  =  mean  temperature  difference  between  the  steam  and  heated 

gases,  degrees  F. 
W  =  weight  of   gases  passing   through  the  superheater  per  boiler 

horse  power  hour. 
c  =  specific  heat  of  the  gases. 

tt  =  temperature  of  the  gases  entering  superheater,  degrees  F. 
t2  =  temperature  of  the  gases  leaving  superheater,  degrees  F. 

Transposing  equation  (23), 

S-We«>-U,  (24) 

Ud 

On  account  of  the  great  variation  in  the  values  given  for  U,  and 
the  difficulty  of  determining  d,  tlf  and  t2  for  different  types  of  super- 
heaters, equation  (24)  is  hardly  applicable  in  practice. 

An  empirical  formula  for  determining  the  extent  of  superheating 
surface  in  connection  with  indirect  superheaters  which  appears  to  con- 
form with  practice  is  given  by  J.  E.  Bell  (Trans.  A.S.M.E.,  May,  1907) : 

_  10  £  XOKX 

~2(T-t)-S' 
in  which 

x  =  square  feet  superheating  surface  per  boiler  horse  power. 
S  =  superheat,  degrees  F. 

T  =  temperature  of  the  products  of  combustion  where  the  super- 
heater is  located. 
t  =  temperature  of  the  saturated  steam. 

The  value  of  T  may  be  found  from  the  equation 
1 


(T  - 
in  which 


=  0.172  H  +  0.294,  (26) 


H  =  the  per  cent  of   boiler-heating  surface  between  the  point  at 

which  the  temperature  is  T  and  the  furnace, 
fas  in  (25). 


SUPERHEATED  STEAM;  SUPERHEATERS 


173 


S3SVO  -IB  U3AO   aass»d  aovjans  ONU.V:JH  uaxvM  Jo  O.N33  H3d. 


174  STEAM  POWER  PLANT  ENGINEERING 

Equation  (26)  is  based  upon  the  assumption  that  the  heat  trans- 
ferred from  the  gases  to  the  water  is  directly  proportional  to  the  differ- 
ence in  temperature;  that  the  furnace  temperature  is  2,500  degrees  F.; 
flue  temperature  500  degrees  F.;  steam  pressure  175  pounds  per  square 
inch  gauge;  one  boiler  horse  power  is  equivalent  to  10  square  feet  of 
water-heating  surface. 

Example:  What  extent  of  heating  surface  is  necessary  to  superheat 
saturated  steam  at  175  pounds  gauge  pressure,  200  degrees  F.,  if  the 
superheater  is  placed  in  the  boiler  setting  where  the  gases  have  already 
traversed  40  per  cent  of  the  water-heating  surface? 

Substitute  H  =  0.4  and  t  =  378  in  equation  (26), 

=  0.172  X  0.4  +  0.294 


x  = 


(T  -  378)0'1 

T  =  950. 

Substitute  T  =  950  and  S  =  200  in  equation  (25), 

10  X  200 

2  (950  -  378)  -  200 

=  2.12  square  feet. 

The  curve  in  Fig.  92a  was  plotted  from  equation  (26),  and  gives  a 
ready  means  of  determining  T  and  of  observing  the  law  governing 
heat  absorption  by  the  boiler  between  furnace  and  breeching.  The 
abscissas  represent  the  temperatures  of  the  hot  gases  at  different  points 
in  their  path  between  furnace  and  breeching.  The  ordinates  represent 
(1)  the  per  cent  of  boiler-heating  surface  passed  over  by  the  hot  gases, 
and  (2)  the  per  cent  of  the  total  heat  generated  which  is  absorbed  by 
this  heating  surface. 

In  the  use  of  equation  (26)  the  probability  of  error  is  greatest  when 
considering  a  point  near  the  furnace,  since  large  quantities  of  heat  are 
transmitted  to  the  tubes  by  radiation  from  the  fuel  bed  which  are  not 
taken  account  of.  For  most  practicable  purposes  the  assumption  is 
sufficiently  accurate. 

For  the  application  of  the  curve  in  Fig.  92a  to  the  design  of  direct 
and  indirect  superheaters  for  various  degrees  of  superheat,  see  "  Stir- 
ling," published  by  the  Stirling  Boiler  Company,  pp.  92-96. 

113.  Performance  of  Superheaters.  —  Published  tests  of  both  directly 
and  indirectly  fired  superheaters  cover  such  a  wide  range  of  conditions 
of  installation  and  operation  that  general  conclusions  cannot  be  drawn, 
but  it  may  be  of  interest  to  note  briefly  the  performances  in  a  few 
specific  cases. 


SUPERHEATED  STEAM;  SUPERHEATERS 


175 


The  curves  in  Figs.  93,  94,  and  95  are  plotted  from  tests  of  a  Babcock 
&  Wilcox  boiler,  with  5000  square  feet  of  water-heating  surface, 
equipped  with  superheating  coils  of  1000  square  feet  area,  as  illustrated 

in  Fig.  62.     The  furnace  with  ordi-  

nary  short  ignition  arch  was  pro- 
vided with  chain  grate  of  75 
square  feet  area. 

Fig.  93  shows  the  relation  be- 
tween degrees  of  superheating  and 
total  horse  power  of  boiler  and 
superheater. 

Fig.  94  shows  the  relation  be- 
tween the  horse  power  produced 


PI  3 
wl 

" 


200  400  600  800 

Horse  Power  Produced  in  Boiler 


FIG.  94. 


900 


700 


500 


400 


300 


200 


100 


Horse  Power  Developed  hi  Superheater 

:,Sg8£gg288isi 

duced  in  the  Superheater  of  that  Devel- 
oped in  the  Boiler. 

1 

•• 

/ 

••/ 

{ 

'••'/ 

1 

/• 

1 

'/..• 

1 

• 

'.'•'  :/ 

•/' 

'/ 

. 

^..'. 

/  • 

./. 

/ 

(•• 

/ 

/ 

/ 

•/ 

/ 

./ 

/ 

/ 

^ 

X 

.  • 

^ 

50          100          150          200 
Degrees  in  Superheat— F 


250 


50  100  150 

Deprees  of  Superheat— F 


200 


FIG.  93.     Relation  of  Degree  of  Superheat 
to  Total  Horse-Power  Developed. 


FIG.  95.    Relation  of  Degree  of  Superheat 
to  Horse-Power  of  Superheater. 


in  the  boiler  and  the  percentage  of  boiler  horse  power  produced  in  the 
superheater. 

Fig.  95  shows  the  relation  between  the  degree  of  superheat  obtained 
and  the  horse  power  developed  in  the  superheater. 


176 


STEAM  POWER   PLANT  ENGINEERING 


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179 


Tables  19  to  21  are  taken  from  the  report  of  Otto  Berner  ("  Zeit.  d. 
Ver.  Deut.  Eng."  and  reprinted  in  Power  August,  1904). 

Table  19  compares  the  heat  efficiency  of  a  steam  plant  equipped  with 
directly  and  with  separately  fired  superheaters,  the  former  showing  a 
much  higher  efficiency. 

Table  20  compares  different  boilers  with  and  without  flue  super- 
heaters, showing  the  effect  upon  the  temperature  of  the  flue  gases. 
The  gain  in  heat  efficiency  of  the  entire  plant  due  to  the  use  of  the  super- 
heater is  decisive  in  each  case. 


TABLE    22. 

(Engineer,  U.  S.,  May  1,  1904.) 


With  Superheater. 

Without   Superheater. 

Tirn6  of  start 

1  2  noon    FVb    8 

UA    Itf       T?f»V»      1  1 

n.   M      TTfiK     1  0 

A.M.,  r  CD.  14 

Hours  run                        

24 

24 

Average  steam  pressure  
Average  water  pressure,  triple  expansion, 
head  in  feet  
Average  water  pressure,  compound,  head 
in  feet 

70.3  lb. 

0.99 

7  10 

79.41b. 
1.05 
7  m 

Average  vacuum  of  suction  for  triple  and 
compound,  inches  of  mercurv  

22  90 

23  21 

Total  head  on  triple,  feet  of  water  

29  05 

29  46 

Total  head  on  compound,  feet  of  water  .  .  . 
Total  double  strokes,  triple  

33.04 
30,557 

33.39 
34  114 

Total  double  strokes,  compound  

35,395 

32  158 

Gallons  pumped  from  piston  displacement, 
total   triple 

2  854  023 

3  186  247 

Gallons  pumped  from  piston  displacement, 
total,  compound  
Gallons  pumped  from  piston  displacement, 
total,  triple  combined  
Gallons,  total,  pumped  as  measured  by  weir 
Per  cent  slip  

2,930,706 

5,784,720 
4,492,680 
22  3 

2,662,682 

5,848,930 
4,549,480 
22  2 

Foot  pounds,  weir  

1,163  815  819 

1  184  983  596 

Total  coal  consumed  

5,015  Ib 

6  410  Ib 

Per  cent  refuse  

23  7 

18  7 

Total  refuse  

1  188 

1  203 

Total  feed  water 

38  399 

t>fl  Qfifl 

Duty  per  100  pounds  coal 

23  206  696 

18  486  483 

Duty  per  1,000  pounds  steam 

30  308  498 

23  253  213 

Per  cent  increase  of  work  per  100  pounds  coal 

Per  cent  increase  of  work  per  1,000  pounds  steam.  . 

Per  cent  saving  in  coal  per  foot  pound  work 

Per  cent  saving  in  feed  water  per  foot  pound  work . . 

Average  temperature  steam  leaving  superheater 

Average  temperature  steam  entering  superheater 

Average  degree  superheat 


..25.5 
..30.2 
,.20.2 

.23.2 

527.4deg.  F. 
.320.1  deg.  F. 
.207.3deg.  F. 


180  STEAM  POWER  PLANT  ENGINEERING 

Table  21  shows  the  gain  in  heat  efficiency  due  to  the  use  of  super- 
heaters in  a  number  of  plants  equipped  with  fire-tube  boilers. 

Table  22  gives  the  results  of  tests  on  one  of  the  return  tubular  boilers 
at  the  Spring  Creek  Pumping  Station  of  the  Brooklyn  Waterworks 
(leb.  9,  1904)  with  and  without  a  superheater.  The  superheater, 
of  the  Foster  type,  was  installed  between  the  rear  wall  of  the  setting 
and  the  tube  sheet. 

113a.  Properties  of  Superheated  Steam. —  The  following  equations 
derived  by  Prof.  Goodenough  of  the  University  of  Illinois  and  based 
upon  the  experiments  of  Knoblauch  and  Jakob,  give  a  comparatively 
simple  method  of  determining  the  various  properties  of  superheated 
steam  if  steam  and  entropy  tables  are  not  available.  The  results  as 
obtained  from  these  equations  agree  substantially  with  Marks'  and 
Davis'  Steam  Tables  and  the  1909  Edition  of  Peabody's  Steam  Tables. 

T  =  absolute  temperature  of  the  superheated  steam,  deg.  F. 
p  =  absolute  steam  pressure,  Ibs.  per  sq.  in. 
A  =  total  heat,  B.T.U.  per  pound. 
u  =  intrinsic  energy,  ft.-lbs. 
n  =  entropy. 
Cp  =  true  specific  heat. 
A  =  T  (0.372  +  0.00005  T)-  p  (1  +  0.00035)  ^  +  882.4, 

in  which  log  C  =  9.42383. 
u  =  T  (202.44  +  0.0389  T)  -  -^  (1  +  0.00025  p)  +  686242, 

in  which  log  C  =  12.20551. 
n  =  0.85657  log  T  +  0.0001  T  -  0.25392  log  p 

-  p  (1  +  0.00035  p)  ^.5  -  0.4300, 
in  which  log  C  =  9.31469. 
Cp  =  0.372  +  0.0001  T  -  p  (I  +  0.00035  p)  ^ 
in  which  log  C  =  9.96790. 

The  mean  specific  heat  may  be  obtained  by  subtracting  the  total  heat 
of  the  saturated  steam  from  that  of  the  superheated  steam  and  dividing 
the  difference  by  the  degree  of  superheat. 

The  specific  volume  may  be  determined  from  Linde's  equation  as 
stated  at  the  bottom  of  page  153. 


CHAPTER  VI. 

COAL  AND  ASH-HANDLING  APPARATUS. 

114.  General.  —  The  cost  of  coal  and  its  delivery  into  the  furnace 
are  usually  the  largest  items  in  the  operating  charges,  hence  large 
central  stations  are  located,  when  practicable,  adjacent  to  a  railway 
line  or  water  front,  to  minimize  the  cost  of  handling  coal  and  ashes. 
Isolated  stations  in  the  business  districts  of  large  cities  are  usually 
unfavorably  situated,  so  that  the  cost  of  handling  coal  and  ashes  is  a  large 
percentage  of  the  total  fuel  cost.     In  large  stations  the  amount  of  fuel 
and  ash  handled  frequently  warrants  the  expense  of  elaborate  conveyor 
systems  which  would  not  be  justified  in  smaller  plants.     In  whatever 
way  coal  is  supplied  provision  should  be  made  for  storing  a  quantity 
sufficient  to  operate  the  plant  for  some  time  in  case  the  supply  is  inter- 
rupted, thereby  guarding  against  an  enforced  shut-down. 

If  adjacent  to  a  railway  line,  a  side  track  must  be  provided  for  switch- 
ing the  cars.  As  bottom-dumping  cars  cannot  be  depended  upon, 
provision  should  be  made  for  unloading  by  hand.  If  coal  is  delivered 
by  water,  clam-shell  drop  buckets  are  ordinarily  used  for  unloading 
the  barges.  If  the  power  house  is  located  at  some  distance  from  the 
railroad  or  water  the  coal  is  generally  hauled  by  teams  in  two  to  five- 
ton  loads. 

115.  Coal  Storage.  —  In  small    stations  the  storage  bins   or  coal 
bunkers  may  usually  be  located  within  the  building,  but  in  larger  plants 
the  quantity  of  coal  consumed  daily  is  frequently  such  that  an  immense 
space  would  be  required  to  furnish  storage  capacity  for  even  a  short 
period  of  time.     For  example,   one  of  the  large  central  stations  in 
Chicago  burns  an  average  of  30  tons  of  Illinois   screenings  per  hour 
throughout  the  year.     Allowing  45  cubic  feet  to  the  ton  this  would 
necessitate  a  space  of  45  x  30  x  24  =  34,800  cubic  feet  to  store  coal  for 
one  day's  operation.     A  ten-days  run  would  require  a  coal  pile  50  feet 
wide,  30  feet  high,  and  232  feet  long.     It  is  a  good  plan,  if  the  location 
and  character  of  the  plant  permit,  to  carry  four  or  five  days'  supply 
within  the  plant  and  provide  a  separate  building  for  the  coal  reserve. 
Such  provision  is  made  in  the  power  plant  of  the  New  York  Edison 
Company,  which  has  a  storage  capacity  of  150,000  tons  in  addition  to 
that  of  the  overhead  bunkers. 

181 


182  STEAM  POWER  PLANT  ENGINEERING 

Exposed  coal  piles  are  objectionable,  because  of  freezing  in  winter,  the 
crust  sometimes  freezing  so  hard  as  to  necessitate  the  use  of  dynamite 
to  break  it;  moreover,  a  slow  depreciation  in  heat  value  takes  place, 
especially  with  bituminous  coal.  This  depreciation  is  more  rapid  in 
warm  weather  and  in  the  tropics.  Stored  coal  is  oftentimes  subject  to 
spontaneous  combustion,  particularly  when  there  is  a  large  content  of 
iron  pyrites. 

Coal  bunkers  or  hoppers  are  ordinarily  placed  on  the  same  level  with 
boiler-room  floor  or  above  the  boiler  setting.  The  former  is  the  cheaper 
as  far  as  first  cost  is  concerned,  but  necessitates  additional  handling  of 
the  fuel  before  it  can  be  fed  to  the  stokers.  In  the  overhead  system 
the  coal  gravitates  to  the  stoker  through  down  spouts.  Overhead 
bunkers  are  usually  found  where  real  estate  is  costly.  They  are  gener- 
ally constructed  of  steel  plates  lined  with  concrete  or  of  reenforced 
concrete.  The  bottoms  slope  at  an  angle  of  35  to  45  degrees  and  empty 
into  the  coal  chutes  or  down  spouts.  Fig.  99  shows  the  general  appear- 
ance of  a  single  overhead  bunker  and  Fig.  441  that  of  a  double  bunker. 
In  some  bunkers  the  floors  are  made  with  very  slight  slopes,  but  it  is  not 
advisable  to  use  a  slope  less  than  the  angle  of  repose  of  the  coal,  as  it  may 
be  necessary  to  shovel  the  coal  over  the  spouts.  Convenience  in  framing 
makes  the  45-degree  slope  the  more  desirable.  Separate  bunkers  for 
each  boiler  are  preferred  to  continuous  bunkers,  since  fire  in  the  coal  is 
more  readily  prevented  from  spreading.  In  the  new  power  house  of 
Swift  &  Co.,  Chicago,  111.,  the  bunkers  are  of  circular  cross  section 
instead  of  rectangular  as  is  the  usual  practice.  The  capacity  of  the 
cylindrical  hopper  is  considerably  less  than  that  of  a  rectangular 
hopper  of  the  same  width,  but  is  much  cheaper  to  construct. 

Ash  bins  are  invariably  lined  with  concrete  or  brickwork,  since  the 
corrosive  action  of  the  ashes  would  soon  destroy  the  bare  iron,  and  are 
usually  located  alongside  the  coal  hopper,  as  in  Figs.  97  and  99,  so  that 
they  may  be  discharged  by  gravity.  The  angle  of  repose  of  most  ashes 
is  approximately  40  degrees,  but  the  45-degree  angle  is  preferred  on 
account  of  convenience  in  construction. 

Coal  Storage  :  Power,  April,  1907,  p.  217,  Aug.,  1899,  p.  3,  Nov.,  1904,  p.  651; 
Eng.  Rec.,  Sept.  23,  1905,  p.  534,  June  1,  1902,  p.  532,  July  4,  1903,  p.  4;  Eng. 
News,  July  11,  1907,  June  5,  1902,  p.  463,  April  3,  1903,  p.  272;  West.  Elec.,  Oct. 
28,  1905,  p.  335;  Trans.  A.S.M.E.,  23-473. 

Coal  Storage  under  Water:  Eng.  News,  Dec.  24, 1908;  Eng.  Min.  Jour.,  Dec.  1,  1904, 
p.  975;  Engineering,  Sept.  4,  1903,  p.  863. 

Calorific  Value  of  Weathered  Coals:  Trans.  A.S.M.E.,  20-333;  Bulletin  No.  17, 
Univ.  of  111.,  Aug.  26,  1907. 

Design  of  Coal  and  Ash  Bins  :  Eng.  News,  July  21,  1904,  p.  62;  Eng.  Rec.,  Sept. 
1,  1900,  p.  201;  Power,  Nov.,  1899,  p.  14,  Nov.  1904,  p.  651;  Elec.  Age,  March, 
1907,  p.  141. 


COAL  AND  ASH-HANDLING  APPARATUS  183 

116.  Coal  Conveyors.  —  Coal  is  carried  to  the  stokers  in  a  variety  of 
ways,  depending  upon  the  location  of  the  plant,  the  type  of  stokers, 
and  the  personal  tastes  of  the  builder.     Of  the  various  methods  the 
following  are  the  most  common  : 

1.  Hand  shoveling  from  coal  pile  to  furnace. 

2.  Wheelbarrow  or  hand  car  and  shovel.    - 

3.  Bucket  conveyor. 

4.  Belt  conveyors. 

5.  Hoist  and  hand  cars. 

6.  Hoist  and  automatic  cable  cars. 

7.  Combinations  of  the  above. 

Coal-Handling  Plants  for  Power  Houses  :  Am.  Elec.,  June,  1900,  p.  266,  Oct., 
1901,  p.  486;  Cassier's,  April,  1905,  p.  480;  Elec.  World.,  Dec.,  1901,  p.  463;  Engr., 
U.S.,  July  1,  1904,  p.  461;  Jan.  1,  1905,  p.  4;  Eng.  Rec.,  April  5,  1902,  p.  322. 

Ash  Handling:  Elec.  World.,  Oct.  5,  1901,  p.  569;  Eng.  News,  Oct.  19,  1905, 
p.  403;  Eng.  Rec.,  May  10,  1902,  p.  435,  Jan.  17,  1903,  p.  85,  Feb.  7,  1903,  p.  153, 
Oct.  28,  1905,  p.  482,  Oct.  7,  1905,  p.  396,  Dec.  9,  1905,  p.  655;  Power,  Oct., 
1904,  p.  507,  July,  1904,  p.  422;  St.  Ry.  Jour.,  Jan.  5,  1901,  p.  11. 

117.  Hand  Shoveling.  —  Where   possible  the  coal  is  dumped  direct 
from  the  cars  or  wagons  into  bins  located  in  front  of  the  boilers.     In 
such  instances  one  man  may  handle  the  coal  and  ashes  and  attend  to 
the  water  level  of  200  horse  power  of  boilers  equipped  with  common 
hand-fired  furnaces.     With  stoking  and  dumping  grates  300  horse  power 
may  be  controlled  by  one  man  and  from  800  to  1000  horse  power  with 
chain-grate   stokers.     This   refers  of  course  to   average  good  coal  not 
too  high  in  ash  nor  productive  of  much  clinker.     Sometimes  the  coal 
cannot  be  stored  in  front  of  the  boilers  but  must  be  hauled  by  wheel- 
barrow, cart,  or  rail  car.     For  distances  over  100  feet  and  quantities 
over  20  tons  per  day  the  cost  of  handling  the  coal  in  this  way  may 
justify  the  installation  of  an  automatic  conveyor  system.     Hand-fired 
furnaces  and  manual  handling  of  coal  and  ashes  are  usually  associated 
with  small  plants  of  500  horse  power  and  under,  but  a  number  of  large 
stations  are  operated  in  this  way  with  apparent  economy.     A  notable 
example  is  the  new  (1907)  steam-power  plant  of  the  Wood  Worsted 
Mill,  Lawrence,  Mass.,  in  which  40  return  tubular  boilers  are  fired  by 
hand.     A  tipcart  with  a  capacity  of  one  ton  brings  the  coal  a  distance 
of  100  to  200  feet  to  the  firing  floor,  and  firemen  shovel  it  on  to  the  grate. 
Four  men  are  stationed  at  the  coal  pile.     One  man  drives  two  carts  (one 
of  which  is  being  filled  while  the  other  is  gone  with  its  load),  sixteen 
firemen  attend  to  the  furnaces,  and  two  men  dispose  of  the  ashes. 


184  STEAM  POWER  PLANT  ENGINEERING 

Most  large  plants,  however,  are  equipped  with  conveying  machinery, 
not  so  much  because  of  the  possible  reduction  in  cost  of  operation, 
taking  into  consideration  all  charges  fixed  and  operating,  as  because 
of  the  large  and  often  unreliable  labor  staff  which  it  dispenses  with. 
Hand  shoveling  is  sometimes  necessary  even  with  modern  unloading 
devices  on  account  of  the  freezing  of  coal  in  the  cars.  This  is  par- 
ticularly true  of  washed  coals,  and  it  is  not  unusual  to  have  an  entire 
car  load  solidly  frozen.  In  this  case  it  has  to  be  picked  and  shoveled 
by  hand,  or  the  unloading  tracks  must  be  equipped  with  steam  pipes 
and  outfits  for  thawing  purposes.  A  good  man  is  capable  of  shoveling 
40  to  50  tons  of  coal  in  eight  hours  when  unloading  a  car,  provided 
it  is  only  necessary  to  shovel  the  coal  overboard. 

118.  Bucket  Conveyors.  —  One  of  the  most  common  methods  of 
automatically  handling  the  coal  from  car  to  bunker  is  by  means  of  an 
endless  chain  of  traveling  buckets.  Many  of  the  largest  central  stations 
in  this  country  are  equipped  with  such  systems.  The  details  of  opera- 
tion are  best  illustrated  by  a  few  examples. 

Fig.  96  gives  a  diagrammatic  arrangement  of  the  link-belt  over- 
lapping pivoted  bucket  carrier,  and  Fig.  97  illustrates  its  application 
to  a  typical  boiler  plant.  Coal  is  discharged  from  the  railway  cars 
into  a  track  hopper  and  from  there  delivered  by  a  "  feeding  apron  " 
into  a  crusher  which  reduces  it  to  such  a  size  as  can  be  conveniently 
handled  by  the  stokers.  It  is  then  discharged  into  a  short  bucket 
conveyor,  which  carries  it  to  the  main  system  of  buckets,  and  it  is 
elevated  to  the  proper  level  and  discharged  into  the  overhead  bunkers. 
The  discharge  is  effected  by  special  tripping  devices  which  engage  the 
buckets  and  turn  them  over.  The  ashes  are  dumped  from  the  ash  pit 
through  a  series  of  chutes  into  the  lower  run  of  buckets,  by  which  they 
are  elevated  and  discharged  into  the  ash  hopper  alongside  the  coal 
bunkers.  From  the  ash  hopper  the  ashes  discharge  by  gravity  directly 
into  the  railway  cars  below.  The  system  is  operated  by  means  of  two 
motors,  one  driving  the  crusher  and  the  other  the  main  bucket  system. 
The  buckets  are  made  of  either  sheet  steel  or  malleable  iron. 

In  Fig.  96  the  coal  is  fed  to  the  crusher  by  the  "  reciprocating  feeder," 
which  is  usually  placed  directly  under  the  track  hopper.  The  feeder 
consists  of  a  heavy  steel  plate  mounted  on  rollers  and  having  a  recip- 
rocating movement  effected  by  a  crank  mechanism  from  the  carrier. 
The  amount  of  coal  delivered  depends  upon  the  distance  the  plate  moves, 
and  this  can  be  varied  by  changing  the  throw  of  the  eccentric.  The 
number  of  strokes  corresponds  to  the  number  of  buckets.  Any  size 
coal  can  be  readily  handled.  When  the  distance  from  track  hopper 
to  carrier  is  so  great  that  the  reciprocating  feeder  is  not  practicable  a 


COAL  AND  ASH-HANDLING  APPARATUS 


185 


186 


STEAM   POWER  PLANT  ENGINEERING 


.2* 

I 


COAL  AND  ASH-HANDLING  APPARATUS 


187 


188 


STEAM   POWER  PLANT  ENGINEERING 


FIG.  99.    Coal  and  Ash-Handling  System  in  the  Power  House  of  the  South  Side 
Elevated  Railway  Company,  Chicago. 


COAL  AND  ASH-HANDLING  APPARATUS  189 

continuous  or  "  belt "  feeder  is  used  to  supply  the  crusher  with  fuel. 
The  "  equalizing  gear  "  is  designed  to  impart  a  pulsating  motion  to  the 
driving  sprocket  wheel  which  will  counteract  the  natural  pulsation  to 
which  long  pitch  chains  are  subject,  producing  violent  increase  of  the 
normal  strain  at  frequent  intervals.  This  is  accomplished  by  driving 
the  spur  wheel  with  an  eccentric  pinion,  causing  the  pitch  line  to  describe 
a  series  of  undulations  corresponding  to  the  number  of  sprockets  on 
the  chain  wheel.  Figs.  99  and  100  show  the  general  arrangement  of 
crusher  and  "  cross  conveyor "  in  the  old  portion  of  the  South  Side 
Elevated  Power  House,  Chicago. 

A  coal  and  ash  system  similar  to  the  one  illustrated  in  Fig.  97  for  a 
plant  consisting  of  eight  350-horse-power  boilers  will  cost  in  the  neigh- 
borhood of  $8,000,  completely  installed.  This  does  not  include  the 
cost  of  coal  and  ash  bunkers. 

The  Hunt  conveyor,  Fig.  101,  while  usually  called  a  "  bucket  " 
conveyor,  is  in  fact  a  series  of  cars  connected  by  a  chain,  each  having 
a  body  hung  on  pivots  and  kept  in  an  upright  position  by  gravity.  The 
chain  is  driven  by  pawls  instead  of  by  sprocket  wheels.  The  "  buckets  " 
are  upright  in  all  positions  of  the  chain,  consequently  the  chain  can  be 
driven  in  any  direction.  The  change  of  direction  of  the  chain  is  accom- 
plished by  guiding  the  carriers  over  curved  tracks.  The  chain  moves 
slowly,  and  the  capacity  is  governed  by  the  size  of  the  buckets.  The 
ordinary  size  buckets  carry  two  cubic  feet  of  coal  and  move  at  a  rate 
of  fifteen  buckets  a  minute,  carrying  about  40  tons  per  hour.  Two 
methods  of  filling  the  buckets  are  employed,  the  "  measuring  "  and  the 
"  spout  filler."  In  the  former  each  bucket  is  separately  filled  with  a 
predetermined  amount  by  a  suitable  "  measuring  feeder."  In  the 
latter  the  material  is  spouted  in  a  continuous  stream,  necessitating 
the  use  of  overlapping  buckets  to  prevent  spilling  of  the  material. 
Fig.  102  shows  an  application  of  the  Hunt'  system  to  the  old  plant 
of  the  Baltimore  United  Railways  and  Electric  Company. 

Fig.  103  gives  a  sectional  elevation  of  the  coal  and  ash-handling 
machinery  at  the  power  plant  of  the  Commercial  National  Bank  Build- 
ing, Chicago.  Underneath  the  sidewalk  on  the  Clark  Street  side  of  the 
building  is  a  coal-storage  bin  of  600  tons  capacity,  served  with  a  bucket 
conveyor.  One  leg  of  the  conveyor  reaches  down  to  a  level  below  the 
track  of  the  Illinois  Tunnel  Company.  By  this  arrangement  coal 
can  be  delivered  either  by  cars  in  the  tunnel  .or  by  wagons  from  the 
street.  In  taking  coal  from  storage  a  gate  at  the  lower  extremity  of 
the  hopper  is  opened  and  the  coal  filling  the  buckets  is  elevated  and 
tripped  into  any  one  of  the  screw  conveyors  leading  from  bucket  con- 
veyor to  boiler  hopper.  The  ashes  are  shoveled  from  the  ash  pits  into 


190 


STEAM  POWER  PLANT  ENGINEERING 


8 


COAL  AND  ASH-HANDLING  APPARATUS  191 


FIG.  101.    Driving  Mechanism  of  Hunt  Conveyor. 


FIG.  102.   Coal  and  Ash-Handling  System  at  the  Old  Power  House  of  the  Baltimore 
United  Railways  and  Electric  Company. 


192 


STEAM   POWER  PLANT  ENGINEERING 


cars  running  in  a  cross  tunnel  under  the  boiler  floor,  and  by  these  cars 
are  transferred  to  a  dump  at  one  side  of  the  boiler  room  and  discharged 
into  Illinois  Tunnel  Company's  cars  for  removal. 


FIG.  103.    Bucket  and  Screw  Conveyor  at  Commercial  National  Bank  Building, 

Chicago,  Illinois. 

119.  Belt  Conveyors.  —  The  Robins  belt  conveyor,  Fig.  104,  consists 
essentially  of  a  thick  belt  of  the  required  width  driven  by  suitable 
pulleys  and  carried  upon  idlers  so  arranged  that  the  belt  becomes 

trough-shaped  in  cross  section. 
The  belt  is  constructed  of  woven 
cotton  duck  covered  with  a 
special  compound  on  the  carry- 
ing side.  The  belt  is  thicker  at 
the  middle  than  at  the  edges, 
since  the  wear  is  greatest  in  a 
line  along  the  center.  The  idlers 


RETURN         IDLERS 


FIG.  104.   Guide  Pulleys,  Robins  Belt 
Conveyor. 


are  carried  by  iron  or  wooden 
framework,  and  are  spaced  from 
3  feet  to  6  feet  between  centers 
on  the  troughing  side  according  to  the  width  of  belt  and  the  weight 
of  the  load.  On  the  return  side  these  distances  range  from  8  to  12  feet. 


COAL  AND  ASH-HANDLING  APPARATUS  193 

High-speed  rotary  brushes  with  interchangeable  steel  bristles  prevent 
wet,  sticky  material  from  clinging  to  the  belt.  Automatic  tripping 
devices  placed  at  the  proper  points  cause  the  material  to  be  discharged 
where  it  is  needed.  The  trippers  consist  essentially  of  two  pulleys, 
one  above  and  slightly  in  advance  of  the  other,  the  belt  running  over 
the  upper  and  under  the  lower  one,  the  course  of  the  belt  resembling 
the  letter  S.  The  material  is  discharged  into  chutes  on  the  first  down- 
ward turn  of  the  belt.  The  trippers  may  be  movable  or  fixed,  single  or 
in  series.  Movable  trippers  are  used  when  it  is  desired  to  discharge 
the  load  evenly  along  the  entire  length,  as,  for  instance,  in  a  continuous 
row  of  bins,  while  fixed  trippers  are  employed  where  the  load  is  to  be 
discharged  at  certain  and  somewhat  separated  points.  The  movable 
trippers  are  made  in  two  forms,  "  hand-driven  "  and  "automatic."  In 
the  former  they  are  moved  from  point  to  point  by  means  of  a  hand 
crank.  The  "  automatic  "  tripper  is  propelled  by  the  conveying  belt 
through  the  medium  of  gearing.  It  reverses  its  direction  automatically 
at  either  end  of  the  run,  and  travels  back  and  forth  continuously  dis- 
tributing its  load.  It  can  be  stopped,  reversed,  or  made  stationary  at 
will.  The  most  notable  installations  of  this  system  are  at  the  96th  Street 
station  and  the  Kingsbridge  station  of  the  Metropolitan  Street  Railway 
Company,  New  York  City. 

120.  Elevating  Tower,  Hand-Car  Distribution.  —  Fig.  105  illus- 
trates the  coal  and  ash-handling  installation  at  the  Aurora  and  Elgin 
Interurban  Railroad  power  house,  Batavia,  111.  Coal  is  delivered  to 
the  plant  by  railroad  cars  which  dump  directly  into  coal  hoppers 
located  inside  a  steel  structure  running  the  entire  length  of  the  building 
and  spanned  by  two  railroad  tracks.  There  are  18  hoppers  constructed 
of  17-inch  brick  walls  fitted  with  steel -plate  bottoms.  Subdividing 
the  storage  space  in  this  manner  makes  it  possible  to  carry  different 
grades  of  coal,  prevents  the  spreading  of  fire,  and  affords  a  simple  con- 
struction for  the  support  of  the  railroad  tracks.  The  basement  of  the 
boiler  room  extends  underneath  the  hoppers,  and  two  lines  of  narrow- 
gauge  tracks  are  imbedded  in  the  concrete  floor.  Turntables  at  the 
center  facilitate  the  switching  of  cars  to  the  elevators  which  rise  through 
the  boiler  room  close  to  the  chimney.  The  cars,  of  one  ton  capacity 
each,  are  of  special  construction,  with  roller-bearing  axles  and  a  com- 
bined ratchet  lift  and  friction  dump.  The  filled  cars  are  pushed  from 
underneath  the  hoppers  to  two  elevators  which  lift  them  to  the  line  of 
tracks  supported  overhead  across  the  boiler  fronts.  They  are  then 
pushed  to  the  hoppers  suspended  above  the  boiler  setting  and  the  coal 
is  dumped.  These  hoppers  have  a  capacity  of  six  tons  each.  From 
the  hoppers  the  coal  is  fed  to  the  stoker  by  an  ordinary  down  spout. 


194 


STEAM  POWER  PLANT  ENGINEERING 


The  ashes  fall  from  the  stokers  into  an  ash  pit,  from  which  they  may  be 
discharged  into  ash  cars.  The  ash  cars  are  elevated  to  a  set  of  tracks 
running  at  right  angles  to  the  main  tracks,  and  are  transferred  to  ash 
bins  located  directly  over  the  coal  bins.  Coal  and  ashes  are  weighed 


GRADE 


TRACK   TO  ELEVATOR 


FIG.  105.    Coal  and  Ash-Handling  System  at  the  Power  House  of  the  Aurora  and  Elgin 
Interurban  Railway,  Batavia,  111. 

in  the  small  cars.  There  are  ten  boilers  in  this  plant  and  four  men  are 
required  to  handle  the  coal  and  ashes.  The  entire  coal  and  ash-handling 
system  cost  about  $10,000,  and  the  cost  of  handling  the  coal  and 
ashes  is  approximately  4  cents  per  ton.  This  does  not  include  wages 
of  firemen  or  water  tenders. 


COAL  AND  ASH-HANDLING  APPARATUS 


195 


131.  Overhead  Storage,  Bucket  Hoist.  —  Fig.  106  gives  a  general 
view  of  the  coal-handling  plant  of  the  Depot  Street  power  house  of  the 
Cincinnati  Traction  Company.  This  installation  is  a  good  example  of 
an  application  of  the  "  overhead  storage  gravity  feed  "  system  to  an 
existing  plant  without  interfering  in  any  way  with  its  operation.  The 
system  consists  essentially  of  a  receiving  pit  below  the  car  tracks  from 
which  the  coal  is  hoisted  to  a  series  of  overhead  bins.  The  coal  storage 
is  outside  the  boiler  house  in  an  independent  structure.  The  bins 
are  of  steel  framework  with  concrete  floors,  and  are  sufficiently  elevated 


PIT     CAPACITY 
SO  TONS 


1-TON  SELF 


Ha. 


Fi-3.  106.    Coal  and  Ash-Handling  System  at  the  Depot  Street  Power  House  of  the 
Cincinnati  Traction  Company. 

to  spout  coal  easily  to  the  stoker  magazine.  The  total  capacity  of  the 
overhead  bins  is  about  1,600  tons.  The  four  bins  or  receiving  pits  have 
a  capacity  of  50  tons  each,  or  approximately  one  car  load,  and  are  so 
situated  that  all  four  may  be  filled  simultaneously  without  shifting  the 
train.  The  coal-handling  apparatus  consists  of  a  one-ton  self-filling 
bucket  operated  on  a  three-motor  electric  crane  running  on  rails  at 
the  top  of  the  storage  bins.  The  coal  is  hoisted  from  the  receiving  pit 
through  suitable  shafts  in  the  bin  structure  and  dumped  into  the  over- 
head hoppers.  The  maximum  capacity  of  the  hoist  is  50  tons  per  hour. 
The  labor  required  to  handle  the  coal  from  car  to  bins  is  performed  by 
one  man  working  five  hours  per  day  and  an  assistant  engaged  a  small 
part  of  the  time  to  dump  cars,  clean  hoppers,  etc.  The  average  daily 
coal  consumption  is  approximately  200  tons.  The  total  cost  of  the 


196  STEAM  POWER  PLANT  ENGINEERING 

equipment  was  about  $18,000  for  the  bins  complete  and  $4,500  for  the 
coal-handling  crane.  The  cost  of  handling  the  coal  and  ashes  is  approxi- 
mately 1.5  cents  per  ton  of  coal.  Including  all  charges  fixed  and  operat- 
ing the  total  cost  of  handling  the  coal  is  about  3.5  cents  per  ton.  This 
does  not  include  wages  of  firemen  or  water  tenders. 

122.  Elevating  Tower,  Cable-Car  Distribution.  —  The  coal  and  ash- 
handling  system  of  the  new  turbine  power  plant  of  the  Detroit  Edison 
Company,  Fig.  107,  is  a  typical  example  of  a  large  station  equipped 
with  elevating  tower  and  cable-car  distributers  instead  of   the  usual 
bucket  conveyor.     The    system    consists    essentially  of    a    lofty  steel 
tower  in  which  are  housed  at  various  levels  a  track  receiving  hopper, 
crushing  rolls  and  screens,  weighing  hopper,  hoisting  apparatus,  etc., 
and  a  small  cable  railway  for  delivery  to  the  bunkers.     The  railroad 
coal  cars  enter  the  tower  on  an  elevated  trestle  18  feet  above  grade, 
below  which  is  a  track  receiving  hopper.     A  two-ton  "  tub  hoist  "  is 
filled  with  coal  from  the  bottom  of  the  receiving  hopper  and  elevated 
to  a  20-ton  bin  at  the  top,  120  feet  above  ground  level.     This  bin  has 
a  grille  bottom  at  one  side  and  under  the  outlet  a  heavy  duty  coal  crusher, 
thus  allowing  the  fine  coal  to  screen  through  directly  while  all  the 
larger  lumps  aie  automatically  delivered  to  the  crusher.     The  hopper 
beneath  this  delivers  to  the  revolving  screen,  which  sorts  the  slack  into 
one  bin  below  and  the  nut  coal  into  the  other.     From  the  two  bins  the 
small  cable  cars  are  filled  for  dumping  into  the  desired  bunkers  over 
the  boiler  rooms.     The  cars  are  arranged  for  automatic   dumping  by 
means  of  adjustable  trips  which  may  be  located   at   any  point.     The 
object  of  separating  the  nut  coal  and  slack  is  to  burn  the  latter  during 
light  or  medium  loads,  keeping  the  former  for  heavy  loads  and  "  peak  " 
overloads.     The  down  spouts  are  double,  with  a  valve  in  each  branch 
operated  from  the  floor,  so  that  either  grade  of  fuel  may  be  drawn  out 
at  any  time  and  in  any  proportion  desired.     The  entire  system  has  a 
capacity  of  from  50  to  75  tons  of  coal  per  hour  and  is  driven  by  steam 
engines,  with  the  exception  of  the  revolving  screen  which  is  motor 
driven.     The  ash -handling  system  consists  of  brick-lined  concrete  hop- 
pers underneath  each  pair  of  stokers  which  discharge  their  contents 
by  gravity  into  the  small  cars  operated  on  the  track  system  in  the 
boiler-house  basement. 

When  handling  275  tons  per  day  of  24  hours  the  cost  of  operation 
is  approximately  12.5  cents  per  ton  from  coal  car  to  ash  car,  including 
wages  of  firemen  and  water  tenders. 

123.  "Vacuum"   Ash   Conveyor.  —  Fig.   108  gives  a  diagrammatic 
arrangement  of   a  recently  patented  ash-conveying  system  depending 
upon  the  velocity  of  a  column  of  air  for  moving  the  ashes.     The  system 


COAL  AND  ASH-HANDLING  APPARATUS 


197 


RE     El V  ING 
BIN 


FIG.  107.   Coal  and  Ash-Handling  System  at  the  Power  House  of  the 
Detroit  Edison  Company, 


198 


STEAM  POWER  PLANT  ENGINEERING 


is  simple  in  operation  and  low  in  first  cost.  One  end  of  special  cast- 
iron  header  F  leads  to  the  ash  pits  of  the  various  boilers  by  means  of 
branch  tubes,  and  the  other  end  is  connected  with  a  sealed  separating 
chamber  A.  Each  branch  pipe  is  fitted  with  simple  circular  openings 
directly  underneath  each  ash-pit  door  for  admitting  ashes  and  which 
are  kept  covered  except  when  in  operation.  Exhauster  E  creates  a 
partial  vacuum  in  chamber  A  and  draws  in  air  at  a  high  velocity  from 
the  opening  in  the  ends  of  the  branch  pipes.  Ashes  raked  into  the 


FIG.  108.   Diagrammatic  Arrangement  of  the  "Vacuum"  Ash-Handling  System. 

pipes  through  the  openings  are  caught  by  the  rapidly  moving  column 
of  air  and  forced  into  chamber  A.  The  ashes  fall  to  the  bottom  and  are 
fed  into  the  main  ash  pit  by  a  slowly  revolving  ash  valve  B.  Air  and 
dust  are  withdrawn  from  the  top  of  the  separator  chamber  through 
pipe  G  and  discharged  to  the  stack  or  to  waste.  A  spray  is  introduced 
into  pipe  F  to  reduce  dust.  The  process  is  a  continuous  one,  and  the 
ashes  may  be  completely  removed  from  the  ash  bin  without  interfering 
with  the  operation  of  the  exhauster.  In  a  later  construction  the  ash 


COAL  AND  ASH-HANDLING  APPARATUS 


199 


,1 


3 


200 


STEAM  POWER  PLANT  ENGINEERING 


COAL  AND  ASH-HANDLING  APPARATUS  201 

bin  and  separating  chamber  are  included  in  one  chamber,  thus  doing 
away  with  the  revolving  ash  valve  and  the  small  motor  operating  it. 
In  this  latter  design  the  bin  is  never  completely  empty,  a  certain  depth 
of  ashes  being  maintained  to  seal  the  bottom  at  all  times. 

At  the  Armour  Glue  Works,  Chicago,  111.,  this  system  is  applied  to  a 
boiler  plant  of  thirteen  boilers,  aggregating  4,800  horse  power,  and  cost, 
completely  installed,  $5,600.  As  originally  installed  the  separating 
chamber  had  a  volume  of  about  35  cubic  feet  and  the  suction  intake  was 
placed  58  feet  above  the  ash-pit  level.  The  revolving  ash  valve  made 
about  13  r.p.m.,  and  was  driven  by  a  one -horse-power  motor.  In  the 
present  installation  the  separating  chamber  and  motor-operated  ash 
valve  are  dispensed  with  and  the  discharge  pipes  lead  directly  into  the 
main  ash  bin,  which  has  a  capacity  of  60,000  pounds  of  wet  ashes  and  is 
constructed  of  five-sixteenths-inch  sheet  iron.  The  exhauster  (a  30-foot 
Root  blower)  has  a  capacity  of  about  8,000  cubic  feet  per  minute  at 
265  r.p.m.,  and  is  driven  by  a  75 -horse-power  motor.  Under  normal 
conditions  of  operation  the  motor  requires  50  horse  power  when  deliver- 
ing 250  pounds  of  ash  per  minute,  and  the  vacuum  on  the  suction  side 
of  the  exhauster  is  3.3  inches  of  mercury.  The  pipe  from  the  ash  bins 
to  the  separating  chamber  is  10  inches  in  diameter  and  is  constructed 
of  extra  heavy  chilled  cast-iron  pipe.  The  piping  from  the  separating 
chamber  to  exhauster  and  to  stack  is  22  inches  in  diameter  and  is  con- 
structed of  number  16  and  number  20  galvanized  iron.  The  ashes  are 
raked  by  hand  from  the  ash  pits  to  the  suction  openings  of  the  branch 
pipes,  and  are  handled  dry,  the  dust  being  taken  along  with  the  ashes. 
Elbows  are  soon  worn  out  by  the  abrasive  action  of  the  ashes,  and  tees 
are  used  instead,  since  the  accumulation  in  the  "  dead  "  end  receives  the 
impact  and  takes  up  the  wear.  The  cost  of  handling  the  ashes  in  this 
installation  is  approximately  7  cents  per  ton. 

124.  Cost  of  Handling  Coal  and  Ashes.  —  In  large  stations  where  a 
number  of  men  are  employed  to  handle  coal  and  ashes  only  it  is  a  simple 
matter  to  divide  the  cost  of  handling  into  the  various  stages,  thus  : 

1.  Cost  of  unloading  cars  or  barges. 

2.  Cost  of  conveying  coal  to  bunkers. 

3.  Cost  of  feeding  coal  to  furnace. 

4.  Cost  of  removing  ashes. 

These  costs  are  usually  expressed  in  cents  or  dollars  per  ton  of  coal 
burned,  or  in  terms  of  cents  or  dollars  per  horse  power  hour  or  kilo- 
watt hour  of  main  prime  mover  output.  Item  number  3  is  oftentimes 
included  under  "  boiler-room  attendance  "  and  items  1,3,  and  4  under 
"  coal  and  ash  handling."  Not  infrequently  all  four  items  are  included 
under  "  attendance."  So  much  depends  upon  the  character  of  stokers 


202  STEAM  POWER  PLANT  ENGINEERING 

and  furnace,  size  of  boilers,  and  the  like,  that  general  figures  on  the  cost 
of  handling  the  coal  and  ashes  are  of  little  value  unless  accompanied  by 
a  description  of  the  equipment.  For  the  sake  of  general  comparison 
the  most  satisfactory  method  of  expressing  the  cost  is  in  dollars  per  ton 
of  coal  from  coal  car  to  ash  car.  This  includes  wages  of  coal  and  ash 
passers,  repair  men,  and  boiler  tenders.  In  small  stations  the  coal 
and  ash  handling  is  done  by  the  boiler  tenders,  in  which  case  it  is 
impracticable  to  separate  the  items  mentioned  above,  and  the  cost  is 
ordinarily  included  under  attendance.  An  average  figure  for  handling 
coal  by  barrow  and  shovel  is  not  far  from  1.6  cents  per  ton  per  yard 
up  -to  the  distance  of  five  yards,  then  about  0. 1  cent  per  ton  per  yard 
for  each  additional  yard.  With  automatic  conveyors  the  operating 
cost,  not  including  wages  of  firemen  and  water  tenders,  varies  with  the 
size  of  plant  and  the  type  of  conveyor,  and  ranges  anywhere  from  a 
fraction  of  a  cent  per  ton  to  four  or  five  cents  per  ton.  The  larger  the 
plant  and  the  greater  the  amount  of  coal  handled  the  lower  will  be  the 
cost  per  ton.  In  comparing  the  relative  costs  of  manual  and  automatic 
handling,  fixed  charges  of  at  least  15  per  cent  of  the  first  cost  of  the 
mechanical  equipment  should  be  charged  against  the  latter  in  addition 
to  the  cost  of  operation.  In  large  central  stations  equipped  with  stokers 
and  conveyors  and  consuming  200  tons  or  more  of  coal  in  twenty-four 
hours,  the  cost  of  handling  the  coal  from  coal  car  to  ash  car,  including 
wages  of  firemen  and  water  tenders,  will  range  between  10  cents  and  18 
cents  a  ton. 

125.  Coal  Hoppers.  —  Fig.  109  shows  a  front  and  side  elevation  of 
a  typical  set  of  stationary  weighing  hoppers  as  applied  to  the  boilers 
of  the  Quincy  Point  power  plant  of  the  Old  Colony  Street  Railway 
Company,  Quincy  Point,  Mass.  Each  battery  of  boilers  is  provided 
with  an  independent  set  of  hoppers.  The  bottoms  of  the  overhead 
coal  bunkers  lead  into  the  small  hoppers  A,  A.  The  operation  of  any 
single  weighing  hopper  is  as  follows:  Coal  is  fed  from  the  overhead 
bunkers  to  weighing  hopper  H  by  means  of  valve  V.  The  weight  of 
coal  in  the  weighing  hopper  is  transmitted  by  a  system  of  levers  and 
knife  edges  to  the  inclosed  scale  beam  /  and  noted  in  the  usual  way. 
The  weighed  charge  of  coal  is  then  admitted  to  the  down  spout  S  by 
means  of  valves  similar  to  those  at  V. 

Although  separate  weighing  hoppers  for  each  battery,  as  illustrated 
in  Fig.  109,  offer  many  advantages,  they  are  quite  costly  and  it  is  not 
unusual  to  install  one  or  more  large  weighing  hoppers  mounted  on 
overhead  traveling  carriages  so  that  one  may  supply  a  number  of 
boilers  (Fig.  110).  At  the  Armour  Glue  Works,  Chicago,  the  coal  supply 
is  stored  in  one  large  overhead  bunker  of  1000  tons  capacity.  A  five- 


COAL  AND  ASH-HAXDLIXG  APPARATUS 


203 


n 


FIG.  109.    Stationary  Coal  Weighing  Hoppers. 


FIG.  110.    Traveling  Coal  Hoppers. 


204 


STEAM  POWER  PLANT  ENGINEERING 


ton  motor-driven  traveling  hopper  receives  its  supply  from  this  central 
bunker  and  delivers  it  to  the  various  boilers.     One  man  operates  the 


FIG.  111.   Common  Slide  Coal  Valve. 


FIG.  112.    Simplex  Coal  Valve. 


traveling  hopper,  tends  to  the  coal  valves,  and  supplies  all  boilers  with 

coal. 

Weighing  hoppers  are  sometimes  made  automatic;  that  is,  the  opening 

and  closing  of  valves,  feeding  of  coal,  and  recording  of  weight  are  auto- 
matically performed  by  the 
weight  of  the  coal  itself.  The 
scale  is  set  for  discharges  of  a 
certain  weight  and  continues 
to  discharge  this  amount  auto- 
matically. In  the  few  plants 
which  are  equipped  with  auto- 
matic weighing  hoppers  the 
capacity  of  the  hopper  is 
approximately  100  pounds  per 
discharge.  These  hoppers  are 
necessarily  more  complicated 
and  more  costly  than  the 
ordinary  weighing  hoppers, 
and  it  is  a  question  whether 


FIG.  113.    Duplex  Coal  Valve. 


the     advantages     offset     the 
extra    first    cost    and    main- 
tenance charges.     A  small  automatic  hopper  of  100  pounds  discharge 
capacity  costs  approximately  $400  as  against  $250  for  the  ordinary 
weighing  device. 


COAL  AND  ASH-HANDLING  APPARATUS 


205 


126.  Coal  Valves.  —  Figs.  Ill  to  115  illustrate  the  principles  of  a  few 
well-known  coal  valves.  They  may  be  conveniently  grouped  into  two 
classes  according  to  the  location  of  the  coal  pocket:  (1)  those  drawing 
the  coal  from  overhead  bunkers  and  (2) 
those  drawing  from  the  side  of  a  bin.  In 
the  first  class  come  the  simple  slide  valve, 
the  simplex  and  duplex  rotating  valve.  In 
the  latter  are  the  flap  valve  and  the 
rotating  valve.  They  are  made  in  various 
sizes  and  designs,  but  those  illustrated  are 
examples  of  the  most  common  types.  The 
simple  slide  valve,  Fig.  Ill,  is  applicable 
only  to  small  size  coal  and  to  small  spouts, 
since  coarse  or  lump  coal  may  get  in  the  way 
and  prevent  proper  closing.  The  simplex 
valve,  Fig.  112,  consists  of  a  rotating  jaw 
actuated  by  a  lever.  There  are  no  rubbing 
surfaces,  and  the  jaws  cut  through  the 
material  without  jamming.  The  duplex 


\ 


FIG.  114.    Common 
Coal  Valve. 


Flap 


FIG.  115.    "  Seaton  "  Coal  Valve. 


valve,  Fig.  113,  consists 
of  two  rotating  jaws  con- 
nected to  a  common 
actuating  lever.  The 
jaws  move  simultane- 
ously, so  that  even  a  par- 
tially open  valve  delivers 
the  coal  centrally.  When 
closing  the  valve  the  flow 
is  gradually  stopped  by 
the  decreasing  width  of 
the  opening  and  there  is 
but  little  resistance  to 
The  largest  valve  can  easily  be  operated 


the  movement  of  the  jaws. 
by  hand. 

The  flap  valve,  Fig.  114,  is  the  simplest  form  for  drawing  coal  from  a 
side  bin.     It  consists  merely  of  an  iron  flap  hinged  to  the  bottom  of 


206  STEAM  POWER  PLANT  ENGINEERING 

the  chute.  The  valve  is  lowered  to  let  the  coal  run  over  its  top  and  is 
raised  to  stop  the  flow.  It  cannot  be  clogged  or  get  jammed  in  closing. 
The  flap  is  raised  and  lowered  by  a  simple  lever.  For  very  large  bins, 
where  the  valves  are  to  be  opened  and  closed  frequently,  the  "  Seaton  " 
valve,  Fig.  115,  is  usually  preferred.  This  valve  consists  of  two  jaws 
EE',  and  TT'  pivoted  to  suitable  framework  at  0  and  actuated  by 
lever  A.  The  valve  is  shown  fully  closed.  Raising  lever  A  causes  the 
cut-off  blade  EE'  to  rotate  about  0  and  permits  the  coal  to  flow 
through  the  space  between  the  edge  of  the  jaw  E  and  the  end  of  the 
chute.  The  rate  of  flow  is  regulated  by  the  width  of  this  opening.  The 
cut-off  blade  does  not  reach  a  stop,  hence  there  is  no  possibility  of  a 
lump  of  coal  getting  in  the  way  and  preventing  the  prompt  closing  of 
the  valve. 

Coal  and  Ash-Handling  Installations  :  Commonwealth  Edison,  Chicago,  Power, 
Dec.,  1906,  p.  718.  Boston  Elevated,  Elec.  World,  Sept.  7,  1901,  p.  396.  Inter- 
borough  Rapid  Transit  Co.,  New  York,  Elec.  World,  Feb.  4,  1905,  p.  264;  Engr., 
U.S.,  May  15,  1904,  p.  337;  Eng.  News,  Jan.  14,  1904,  p.  41.  Waterside  Station, 
New  York,  Edison  Co.,  Eng.  Rec.,  Sept.  9,  1905,  p.  287.  Detroit  Edison  Co.,  Eng. 
Rec.,  Oct.,  1905,  p.  396.  Brooklyn  Rapid  Transit  Co.,  St.  Ry.  Jour.,  Sept.  23, 
1905,  p.  435.  N.  Y.  C.  and  H.  R.  R.,  St.  Ry.  Jour.,  Nov.  11,  1905,  p.  876.  Aurora, 
Elgin  and  Chicago  Ry.,  Eng.  Rec.,  Feb.  7,  1903,  p.  153.  Missouri  River  Power 
Co.,  Eng.  News,  Oct.  19,  1905,  p.  403.  Brooklyn  Edison  Co.,  Gold  St.  Station, 
Elec.  World,  June  15,  1907. 

Hoisting  and  Conveying  Machinery:    Pro.  A.S.M.E.,  June,  1908. 


CHAPTER  VII. 

CHIMNEYS. 

127.  Chimney  Draft.  — Draft  produced  by  a  chimney  depends  upon 
so  many  conditions  and  involves  such  a  large  number  of  variables  that 
empirical  methods  of  proportioning,  based  upon  actual  performances, 
are  more  to  be  relied  upon  than  theoretical  calculations.  Draft  is 
due  to  the  difference  in  the  weight  of  the  column  of  hot  light  gases 
in  the  stack  and  that  of  the  cooler  and  heavier  surrounding  atmos- 
phere, the  latter  tending  to  flow  into  the  base  and  thereby  force  the 
lighter  gases  out  the  top  of  the  stack.  The  commonly  accepted  theory 
of  chimney  draft  is  based  upon  Peclet's  hypothesis  that  the  flow  through 
the  furnace  flues  and  chimney  may  be  represented  by  the  equation 

(27) 

in  which 

h  =  the  head  of  fluid  producing  the  flow,  feet. 

u  =  velocity  of  the  gases  in  the  chimney,  feet  per  second. 

G  =  a  coefficient  to  represent   the  resistance  to  the   passage  of  air 

through  the  coal. 

I  =  total  length  of  the  path  of  the  gases,  feet. 
?7i  =  area  of  cross  section  divided  by  the  perimeter. 
/   =  a  coefficient  depending  upon  the  nature  of  the  surfaces  over 
which  the  gases  pass. 

From  experiments  on  chimneys  and  boilers  Peclet  gives  in  connection 
with  this  theory  the  following  values  of  coefficients  G  and  /  : 

G  =  12,  /  =  0.012, 

on  the  basis  of  20  to  24  pounds  of  coal  burned  per  square  foot  of  grate 
surface  per  hour.  On  account  of  the  variation  in  practice  of  the  factors 
u,  /,  and  G  and  the  difficulty  of  determining  them  engineers  prefer  to 
use  the  modified  formulas  given  further  on. 

The  difference  of  pressure,  or  intensity  of  draft  may  be  expressed 
theoretically,  ignoring  friction,  as  follows  : 

207 


208  STEAM  POWER  PLANT  ENGINEERING 

Let      H  =  height  of  chimney  in  feet. 

T  =  absolute  temperature  of  the  freezing  point,  degrees  F. 
!Tl=  absolute  temperature  of  the  gases  in  the  chimney. 
T2=  absolute  temperature  of  the  outside  air. 
P  =  average  atmospheric  pressure. 
P2=  observed  atmospheric  pressure. 

W  =  weight  of  a  cubic  foot  of  air  at  32  degrees  F.  and  pressure  P. 
Wl=  weight  of  a  cubic  foot  of  chimney  gas  at  32  degrees  F. 
and  pressure  P. 

Then  the  weight  of  a  cubic  foot  of  hot  gas  in  the  chimney  will  be 


and  the  weight  of  a  cubic  foot  of  cold  air  outside  will  be 

P        T 


The  weight  of  a  column  of  hot  gas  H  feet  high  and  one  foot  square 
will  be 

W,H.       -.  (30) 


Similarly  the  weight  of  the  cold-air  column  will  be 


and  the  difference  in  pressure  or  the  intensity  of  draft  will  be 

(32) 


where  D  is  in  pounds  per  square  foot. 

By  making  -P  =  P2  =  14.7,  T  =  493,  W  =  0.0807,  W,  =  0.084, 
and  D!  =  pressure  in  inches  of  water  (Dt  =  0.192  D),  equation  (32) 
assumes  the  familiar  form 

(33) 


By  assuming  W  =  Wl  =  0.081    and    P  =  14.7,  equation  (32)  may 
be  written 

(34) 


This  latter  form  is  ordinarily  used  where  the  atmospheric  pressure 
differs  considerably  from  that  at  sea  level,  as  at  high  altitudes.  Table  23 
gives  the  density  of  air  and  chimney  gases  at  various  temperatures. 


CHIMNEYS 


209 


Example  :  Required  the  maximum  theoretical  draft  obtainable  from 
a  chimney  150  feet  high,  atmospheric  pressure  14.7  pounds  per  square 
inch,  temperature  outside  air  60  degrees  F.,  temperature  chimney 
gases  550  degrees  F. 

Here      H  =  150,     T2  =  461  +  60  =  521,    7\  =  461  +  550  =1011. 

Substituting  these  values  in  equation  (34), 


=  15°(~521  ~  KHl 


which  is  about  25  per  cent  greater  than  the  draft  actually  obtained, 
and  represents  the  maximum  possible  under  the  given  conditions, 
neglecting  the  resistance  offered  by  the  chimney  and  the  pressure 

TABLE  23. 

DENSITY  AND  SPECIFIC  VOLUME  OF  AIR  AND  CHIMNEY  GASES  AT 
VARIOUS  TEMPERATURES. 


Air. 

Chimney  Gases. 

t 

4 

V 

d 

t 

d 

t 

d 

t 

d 

0 

11.581 

.935 

.086353 

200 

.06334 

430 

.04695 

660 

.03730 

5 

11.706 

.945 

.085424 

210 

.06239 

440 

.04643 

670 

.03697 

10 

11.832 

.955 

.084513 

220 

.06147 

450 

.04592 

680 

.03665 

15 

11.931 

.965 

.083623 

230 

.06058 

460 

.04542 

690 

.03633 

20 

12.085 

.976 

.082750 

240 

.05971 

470 

.04493 

700 

.03602 

25 

12.211 

.986 

.081895 

250 

.05887 

480 

.04445 

710 

.03571 

30 

12.337 

.996 

.081058 

260 

.05805 

490 

.04398 

720 

.03540 

32 

12.387 

1.000 

.080728 

270 

.05726 

500 

.04353 

730 

.03511 

35 

12.463 

1.006 

.080238 

280 

.05648 

510 

.04308 

740 

.03481 

40 

12.589 

1.016 

.079434 

290 

.05573 

520 

.04264 

750 

.03453 

45 

12.715 

1.026 

.078646 

300 

.05499 

530 

.04221 

760 

.03424 

50 

12.841 

1.037 

.077874 

310 

.05428 

540 

.04178 

770 

.03396 

55 

12.967 

1.047 

.077117 

320 

.05358 

550 

.04137 

780 

.03369 

60 

13.093 

1.057 

.076374 

330 

.05290 

560 

.04096 

790 

.03342 

62 

13.144 

1.061 

.076081 

340 

.05224 

570 

.04056 

800 

.03316 

65 

13.220 

1.067 

.075645 

350 

.05159 

580 

.04017 

900 

.03072 

70 

13.346 

1.077 

.074930 

360 

.05096 

590 

.03979 

1000 

.02861 

75 

13.472 

1.087 

.074229 

370 

.05035 

600 

.03942 

1100 

.02678 

80 

13.598 

1.098 

.073541 

380 

.04975 

610 

.03905 

1200 

.02516 

85 

13.724 

1.108 

.072865 

390 

.04916 

620 

.03869 

1300 

.02373 

90 

13.851 

1.118 

.072201 

400 

.  04859 

630 

.03833 

1400 

.02245 

95 

13.976 

1.128 

.071550 

410 

.04803 

640 

.03798 

1500 

.02131 

100 

14.102 

1.138 

.070910 

420 

.04749 

650 

.03764 

1800 

.01848 

110 

14.354 

1.159 

.069665 

2000 

.01698 

d  =  density,  pounds  per  cubic  foot. 

t  =  temperature,  degrees  F. 

s  =  specific  volume,  cubic  feet  per  pound. 

v  =  comparative  volume,  volume  at  32°  =  1. 

Density  of  chimney  gas  taken  0.085  pound  per  cubic  foot  at  32°  F.  and  29.92 
inches  of  mercury. 

(Rankine, "  Steam  Engine,"  gives  the  density  at  32°  F.  as  varying  from  0.084  to 
0.087.) 


210 


STEAM  POWER  PLANT  ENGINEERING 


TABLE  24. 

THEORETICAL    DRAFT    PRESSURE    IN    INCHES    OF  WATER.     CHIMNEY 
100  FEET    HIGH.1 


Temp, 
in  the 
Chim- 
ney. 

Temperature  of  the  External  Air  —  Barometer,  14.7  Pounds  per  Square  Inch.2 

0° 

10° 

20° 

30° 

40° 

50° 

60° 

70° 

80° 

90° 

100° 

200 

.453 

.419 

.384 

.353 

.321 

.292 

.263 

.234 

.209 

.182 

.157 

220 

.488 

.453 

.419 

.388 

.355 

.326 

.298 

.269 

.244 

.217 

.192 

240 

.520 

.488 

.451 

.421 

.388 

.359 

.330 

.301 

.276 

.250 

.225 

260 

.555 

.528 

.484 

.453 

.420 

.392 

.363 

.334 

.309 

.282 

.257 

280 

.584 

.549 

.515 

.482 

.451 

.422 

.394 

.365 

.340 

.313 

.288 

300 

.611 

.576 

.541 

.511 

.478 

.449 

.420 

.392 

.367 

.340 

.315 

320 

.637 

.603 

.568 

.538 

.505 

.476 

.447 

.419 

.394 

.367 

.342 

340 

.662 

.638 

.593 

.563 

.530 

.501 

.472 

.443 

.419 

.392 

.367 

360 

.687 

.653 

.618 

.588 

.555 

.526 

.497 

.468 

.444 

.417 

.392 

380 

.710 

.676 

.641 

.611 

.578 

.549 

.520 

.492 

.467 

.440 

.415 

400 

.732 

.697 

.662 

.632 

.598 

.570 

.541 

.513 

.488 

.461 

.436 

420 

.753 

.718 

.684 

.653 

.620 

.591 

.563 

.534 

.509 

.482 

.457 

440 

.774 

.739 

.705 

.674 

.641 

.612 

.584 

.555 

.530 

.503 

.478 

460 

.793 

.758 

.724 

.694 

.660 

.632 

.603 

.574 

.549 

.522 

.497 

480 

.810 

.776 

.741 

.710 

.678 

.649 

.620 

.591 

.566 

.540 

.515 

500 

.829 

.791 

.760 

.730 

.697 

.669 

.639 

.610 

.586 

.559 

.534 

550 

.863 

.828 

.795 

.762 

.731 

.700 

.671 

.644 

.618 

.593 

.585 

600 

.908 

.873 

.839 

.807 

.776 

.746 

.717 

.690 

.663 

.638 

.613 

1.  For  any  other  height  multiply  the  tabular  figure  by  r^,    where  H  is  the  height  in  feet. 

p 

2.  For  any  other  pressure  multiply  the  tabular  figure  by  ,  where  P  is  the  barometric  pres- 
sure in  pounds  per  square  inch. 

required  to  impart  velocity  to  the  gases.  Table  24  has  been  computed 
from  formula  (34),  and  gives  the  maximum  theoretical  draft  in  a  chim- 
ney 100  feet  high  for  different  flue-gas  temperatures. 

The  intensity  of  draft  required  to  produce  best  results  depends  upon 
the  kind  and  condition  of  fuel,  the  thickness  of  fire,  character  of  grate, 
and  resistance  of  the  breeching,  tubes,  baffles,  dampers,  etc.  As  stated 
above,  the  loss  of  draft  in  the  chimney  proper  approximates  20  per  cent  of 
the  total,  that  in  the  breeching  is  taken  as  0.1  inch  per  100  feet  of  flue, 
and  0.05  inch  for  each  right-angle  bend;  the  loss  in  the  boiler  varies  from 
0.3  to  0.6  inch,  depending  upon  the  type;*  the  loss  in  the  furnace  varies 
between  wide  limits,  and  depends  upon  the  kind  of  fuel  and  the  rate  of 
combustion.  The  curves  in  Fig.  116  compiled  by  the  Stirling  Company 
and  published  in  their  book  "  Stirling  "  give  the  furnace  drafts  necessary  to 
burn  various  kinds  of  fuels  at  different  combustion  rates,  and  give  an  idea 
of  the  influence  of  the  character  of  the  fuel  and  the  rate  cf  combustion. 
*  Specific  figures  may  be  obtained  from  the  manufacturers. 


CHIMNEYS 


211 


5        8        «        -.        §        § 

(U31VM  JO  83HONI)   Uld   HSV 


OOO  OOOOOO 

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QNV  aovNunj  N33Mxaa  aaainSaa  Xdvua  JO 


212  STEAM  POWER  PLANT  ENGINEERING 

Example  :  Determine  the  probable  draft  necessary  to  burn  30  pounds 
bituminous  run  of  mine  per  hour  per  square  foot  of  grate  when  the  out- 
side air  is  60  degrees  F.,  the  temperature  of  the  chimney  gases  550 
degrees,  and  the  flue  is  100  feet  long,  with  two  right-angle  bends. 

The  losses  will  be  divided  approximately  as  follows  : 

Inch. 
Loss  in  furnace  (from  curves  in  Fig.  116)  ....................  0.  17 

Loss  in  boiler  (average)  ...................................  0.  40 

Loss  in  flue,  100  feet  at  0.10  per  100  ........................  0  .  10 

Loss  in  turns,  2  X  0.05  .....................................  0.10 

6T77 

Since  the  loss  in  the  chimney  alone  approximates  20  per  cent  of  the 
total,  0.77  -*•  .80  =  0.96  will  be  the  theoretical  draft  necessary.  From 
equation  (33), 


Substituting  for  the  given  values  of  Z>1;  7\,  and  T2  in  above  equation, 
096-    W7'64       7'95\. 

d\52i      ion) 

From  which, 

H  =  142,  height  of  stack  necessary  to  produce  a  draft  of  0.17  inch 
in  the  furnace. 

Table  25  gives  the  results  of  a  test  of  a  100-foot  unlined  steel  chimney, 
showing  the  variation  in  draft  at  different  points  in  the  stack. 

Theory  of  Chimney  Draft  :  Power,  Oct.,  1896,  p.  18,  Dec.,  1898,  p.  20,  March, 
1906,  Feb.,  1900,  p.  12;  Engr.  U.S.,  Jan.  15,  1903,  May  15,  1902,  p.  313;  Trans. 
A.S.M.E.,  11-451,  762,  772,  974,  984;  Elec.  Rev.,  Lond.,  Oct.  14,  1904. 

138.  Chimney  Formulas.  —  Rational  methods  of  determining  the 
height  and  area  of  chimneys  being  cumbersome  and  unwieldy  and  of 
doubtful  value  for  practical  use,  the  various  empirical  formulas  outlined 
in  Table  26  are  quite  commonly  used.  They  give  good  results  within 
the  limits  of  the  assumptions  upon  which  they  are  based,  but  otherwise 
may  lead  to  absurd  results,  their  applicability  depending  largely  upon 
the  available  data  covering  the  various  losses  with  the  particular  kind, 
quality,  and  condition  of  coal,  and  conditions  of  operation.  Occasionally 
practical  and  local  considerations  fix  the  height  of  the  stack  irrespective 
of  theoretical  deductions.  The  logical  procedure  is  to  determine  first 
the  height  of  chimney  necessary  to  produce  the  draft  at  the  desired  maximum 
rate  of  combustion,  and  then  to  proportion  the  area  by  such  formulas  as 
(2),  (4),  or  (5),  to  suit  the  quantity  of  fuel  to  be  burned. 


CHIMNEYS 


213 


The  following  heights  have  been  found  to  give  good  results  in  plants 
of  moderate  size: 

Feet. 

With  free-burning  bituminous  coal 80 

With  anthracite,  medium  and  large  sizes 100 

With  slow-burning  bituminous 120 

With  anthracite  pea 130 

With  anthracite  buckwheat 150 

With  anthracite  slack. .  175 


TABLE  25. 
CHIMNEY    DRAFT. 

Test  of  a  100-Foot  Unlined  Steel  Chimney  3  Feet  in  Diameter  at  Massachusetts 
Institute  of  Technology.     (Peabody  &  Miller,  "Steam  Boilers,"  p.  121.) 


Draft,  Inches 
of  Water. 

Temperature,  Fah- 
renheit. 

Maximum. 

Minimum. 

Maximum  . 

Minimum. 

Over  the  grate         

0.24 
0.382 

0.410 
0.354 
0.572 
0.440 
0.334 
0.216 
0.122 

0.218 
0.372 

0.374 
0.334 
0.543 
0.414 
0.312 
0.168 
0.086 

At  the  bridge  wall        

Half-way  between  bridge  and  back  end 
of  boiler            

At  the  back  end  of  boiler  

In  uptake  near  boiler  
In  stack  34  feet  above  grate  

403 
396 
380 
370 
345 

389 
374 
368 
354 
314 

In  stack  51  feet  above  grate  

In  stack  68  feet  above  grate  

In  stack  85  feet  above  grate  

The  chimney  serves  two  80-horse-power  boilers.  During  test  one  was 
banked  and  the  combustion  at  the  grate  of  the  working  boiler  was  19.8 
pounds  per  square  foot  of  grate  surface  per  hour.  Coal  burned  per 
hour  590  pounds. 

For  plants  of  800  horse  power  or  more  the  height  of  stack  should 
never  be  less  than  150  feet,  regardless  of  the  kind  of  coal  used. 

Referring  to  Table  26,  formulas  (1),  (2),  (6), (7),  and  (9)  are  based  upon 
a  fuel  consumption  of  13  to  15  pounds  of  anthracite  and  22  to  26 
pounds  of  bituminous  coal  per  square  foot  of  grate  area  per  hour.  In 
formulas  (3),  (4)  and  (9),  the  diameter  is  dependent  solely  upon  the 
quantity  of  coal  burned  per  hour  and  the  height  is  determined  mainly 
by  the  rate  of  combustion  per  square  foot  of  grate.  The  results  accord 
well  with  practice.  With  western  coals  formula  (3)  gives  results  rather 
too  large  and  the  constant  should  be  120  instead  of  180.  Formula  (5)  is 


214  STEAM  POWER  PLANT  ENGINEERING 

perhaps  the  most  used  and  has  met  with  much  approval.     It  is  based  on 
the  assumptions  that 

1.  The  draft  of  the  chimney  varies  as  the  square  root  of  the  height. 

2.  The  retardation  of  the  ascending  gases  by  friction  may  be  con- 
sidered due  to  a  diminution  of  the  area  of  the  chimney  or  to  a  lining  of 
the  chimney  by  a  layer  of  gas  which  has  no  velocity  and  the  thickness 
of  which  is  assumed  to  be  2  inches.     Thus,  for  square  chimneys, 


(35) 

and  for  round  chimneys, 


E  =       O*  -        .     =  A-  0.591  vQ.  (36) 


For  simplifying  calculations  the    coefficient  of  \/  A  may  be  taken  as 
0.6  for  both  square  and  round  chimneys,  and  the  formula  becomes 

E  =  A  -  0.6  VA.  (37) 

3.  The  horse-power  capacity  varies  as  the  effective  area  E. 

4.  A  chimney  should  be  proportioned  so  as  to  be  capable  of  giving 
sufficient  draft  to  permit  the  boiler  to  develop  much  more  than  its  rated 
power  in  case  of  emergencies  or  to  permit  the  combustion  of  5  pounds 
of  fuel  per  rated  horse  power  per  hour. 

5.  Since  the  power  of  the  chimney  varies  directly  as  the  effective 
area  E  and  as  the  square  root  of  the  height  H,  the  formula  for  horse 
power  for  a  given  size  of  chimney  will  take  the  form 

H.P.  =  CE  V~H,  (38) 

in  which  C  is  a  constant,  found  by  Mr.  Kent  to  be  3.33,  obtained  by 
plotting  the  results  from  numerous  examples  in  practice. 
The  formula  then  assumes  the  form 

H.P.  =  3.33  E  V~H  (39) 

or  H.P.  =  3.33  (A  -  0.6  VA)  V~H,  (40) 

from  which 

/HQ   TT  T>  \  2 

(41) 

Table  27  has  been  computed  from  equation  5,  Table  26. 
Many  engineers  simply  adopt  the  following  proportions  : 
Internal  area  of  chimney  at  top,  one-seventh  grate  area  for  bitumi- 
nous coal. 


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CHIMNEYS  21 7 

Internal  area  of  chimney  at  top,  one-ninth  grate  area  for  anthracite 
coal. 

Example :  Determine  the  area  and  diameter  of  a  stack  for  a  2000- 
horse-power  plant  to  operate  under  the  following  conditions:  Rated 
load  2000  horse  power;  maximum  overload  40  per  cent  of  rated;  flue 
150  feet  long,  with  one  right-angle  bend;  average  rate  of  combustion  20 
pounds  of  bituminous  coal  per  square  foot  of  grate  surface  per  hour; 
atmospheric  temperature  60  degrees  F.;  flue-gas  temperature  at  over- 
load 600  degrees  F.;  coal  burned  per  boiler  horse  power,  4  pounds. 

With  modern  types  of  steam  engines  or  turbines  an  overload  of  40 
per  cent  has  little  effect  on  the  economy  of  the  prime  mover,  and  the 
boiler  efficiency  is  but  slightly  reduced,  but  an  additional  allowance  of 
25  per  cent  should  be  made  in  estimating  the  overload  combustion  rate. 

The  maximum  rate  of  combustion  then  will  be 


pounds  per  square  foot  of  grate  surface  per  hour. 

The  draft  required  at  the  point  where  the  flue  enters  the  chimney, 
considering  the  various  losses,  will  be  found  as  follows  : 

Inch. 
Furnace  (see  curves,  Fig.  116)  ....................  .  .........  0.3 

Boiler  ............................  .  ........................  0.4 

Flue,  150  feet  at  0.1  inch  per  100  feet  .....  .................  0.  15 

Turns,  1  at  0.05  ...........................................  0.05 

0.9 

From  formula  (33), 


Substituting  the  following  values  : 

T2  =  60  +  461  =  521  ;       Tl  =  600  +  461  =  1061 

0  9 

Dl  =  theoretical  draft  =7^  =1.12  inch. 

U.o 

.64      7.95 


whence  the  necessary  height  of  stack  is 

H  =  160  feet  (approximately). 


218  STEAM  POWER  PLANT  ENGINEERING 

Substituting  the  value  of  H  in  Kent's  formula,  the  effective  area  is 
found  to  be 

„      0.3  H.P.      0.3  X  2000       ..  _  , 

E  =  — 7= —  = 7= =  47.5  square  feet. 

VH  Vl60 

corresponding  to  an  actual  diameter  of  93  inches. 

Chimney  Design:  Engr.  U.S.,  Jan.  1,  1907,  p.  81,  Feb.  1,  1907,  p.  174;  Am. 
Elecn.,  March,  1904,  p.  123;  Eng.  Rec.,  April  18,  1900,  July  27,  1901,  pp.  52,  82, 
Sept.  21,  1901,  p.  271,  March  1,  1902,  p.  195,  July  19,  1902,  p.  60,  Nov.  22,  1902, 
p.  495,  May  5,  1906,  p.  549;  Power,  Jan.,  1902,  p.  12,  Nov.,  1902,  p.  29,  Dec.  1, 
1903,  p.  66,  Dec.,  1901,  p.  570,  Dec.,  1905,  p.  745;  Am.  Elecn.,  Dec.,  1901,  p.  570; 
Cassier's  Mag.,  July,  1904,  p.  341,  Feb.,  1906,  p.  267;  Engr.  U.S.,  Oct.  1,  1899, 
p.  222,  Sept.  1,  1904,  p.  591,  June  15,  1905,  p.  403;  Eng.  Mag.,  Oct.,  1899,  p.  124, 
Trans.  A.S.M.E.,  11-451;  Eng.  News,  July  20,  1905,  p.  64;  Engr.  U.S.,  Jan.  1, 
1907,  p.  91. 

129.  Height   of   Chimneys    for   Boilers    using    Oil    Fuel.  —  Experi- 
mental data  relative  to  chimneys  for  boilers  using  oil  fuel  are  rather 
meager,  and  discordant,  but  a  study  of  a  number  of  recent  installations 
seems  to  indicate  that  the  area  need  not  exceed  50  per  cent  of  that 
required  by  the  same  boiler  using  bituminous  coal.     A  height  80  to 
90  feet  above  the  grate  usually  affords  sufficient  draft  to  force  the 
boilers  50  per  cent  above  rating,  but  in  a  number  of  large  installa- 
tions the  chimneys  have  been  designed  on  the  coal-burning  basis  so  as 
to  provide  sufficient  capacity  in  case  it  proves  necessary  at  a  future 
date  to  revert  to  the  use  of  coal. 

130.  Classification  of  Chimneys.  —  Chimneys  may  be  grouped  into 
three  classes  according  to  the  material  of  construction: 

1.  Steel. 

2.  Reenforced  concrete. 

3.  Masonry. 

Steel  chimneys  have  many  advantages  and  are  finding  much  favor 
in  large  power  plants,  especially  where  economy  of  space  warrants 
the  erection  of  the  stack  over  the  boiler,  in  which  case  the  structural 
work  of  the  boiler  setting  answers  for  both  boiler  and  chimney. 
Among  the  advantages  are  (1)  ease  and  rapidity  of  construction; 
(2)  less  weight  for  a  given  internal  diameter  and  height;  (3)  less  sur- 
face exposed  to  the  wind;  (4)  lower  cost;  (5)  smaller  space  required; 
(6)  slightly  higher  efficiency  if  properly  calked,  for  there  can  be  no 
infiltration  of  cold  air  as  is  likely  through  the  cracks  in  masonry.  The 
chief  disadvantage  is  the  cost  of  keeping  the  stack  well  painted  to 
prevent  rust  and  the  corrosive  action  of  the  sulphur  in  the  coal. 

Steel  chimneys  may  be 

1.  Guyed. 

2.  Self-sustained. 


CHIMNEYS 


219 


131.  Guyed    Chimneys.  —  Guyed    sheet-iron   or   steel    chimneys  or 
stacks  held  in  position  by  guy  wires  are  employed  in  small  sizes  on 
account  of  their  relative  cheapness.     They  seldom  exceed  52  inches  in 
diameter  and  75  feet  in  height.     A  heavy  foundation  is  unnecessary, 
and  the  stack  may  be  supported  by  the  boiler  breeching.     The  small 
short  stacks  are  ordinarily  riveted  in  the  shop,   ready   for  erection, 
larger  sizes  being  shipped  in  sections  and  riveted  at  the  place  of  instal- 
lation.    The  guy  wires  are  usually  fastened  to  an  angle  iron  or  band  at 
about  two-thirds  the  height,   and  anchored  at  a  distance  from  the 
base  equal  to  the  height  of  the  band  above  the  ground. 

For  very  tall  stacks  two  sets  of  guys  are  used,  from  four  to  six 
wires  being  fastened  to  each  band,  and  designed  to  withstand  a  wind 
pressure  of  30  pounds  per  square  foot  of  projected  area  of  the  stack. 
Turn-buckles  are  employed  to  equalize  tautness.  Table  28  gives  the 
thickness  of  material,  with  approximate  cost  and  weight,  of  guyed 
stacks  of  different  heights  and  areas. 

132.  Self-Sustaining    Steel    Chimneys.  —  Steel    chimneys    over    52 
inches  in  diameter  are  usually  self-supporting.     They  may  be  built 
with  or  without  a  brick  lining,  but  the  lining  is  preferred,  since  it 
prevents  radiation  and  protects  the  inside  from  the  corrosive  action 
of  the  flue  gases.     Since  the  lining  plays  no  part  in  the  strength  of  the 
chimney,  it  is  made  only  thick  enough  to  support  its  own  weight,  and 
usually  of  a  low-grade  fire  brick  or  carefully  burned  common  brick  or 
both.     In  average  practice  the  fire  brick  extends  20  or  30  feet  above 

TABLE   28. 
APPROXIMATE  WEIGHT  AND   COST   OF   GUYED   SHEET-STEEL   CHIMNEYS. 


Height,  Feet. 

Diameter,  Inches. 

Thickness  of  Shell, 
B.W.G. 

Approximate  Weight 
per  Foot,  Pounds. 

40 

18 

16 

13 

45 

20 

16 

14 

45 

22 

14,  16 

20,  15 

50 

24 

14,  16 

22,  16 

50 

26 

14 

23.5 

55 

28 

14 

25 

60 

30 

12,  14 

34,  27 

65 

32 

12,  14 

36,  28 

70 

34 

10,  12 

48,  39 

75 

36 

10,  12 

51,  41 

Approximate  cost  per  pound,  3.5  cents  to  6.5  cents,  including  cost  of  sections 
riveted  and  punched,  ready  for  assembling,  the  higher  figure  referring  to  the 
smaller  stacks. 


220  STEAM  POWER  PLANT  ENGINEERING 

the  breeching,  the  remainder  of  the  lining  being  of  common  brick.  In 
chimneys  up  to  80  inches  internal  diameter,  the  upper  course  is  4J 
inches  thick,  and  increases  4J  inches  in  thickness  for  each  30  to  40  feet 
to  the  bottom.  In  larger  chimneys  about  8  inches  is  the  minimum 
thickness.  The  lining  is  generally  set  in  contact  with  or  close  to  the 
shell,  though  a  space  of  1  to  2  inches  is  sometimes  left  between  the 
brickwork  and  the  shell  to  allow  for  expansion.  This  space  is  occa- 
sionally filled  with  sand. 

Self-sustaining  stacks  may  be  straight  or  tapered,  and  are  generally 
made  with  a  flared  or  bell-shaped  base  whose  diameter  and  length  are 
1J  to  2  times  the  internal  diameter  of  the  stack.  The  base  is  riveted 
to  a  heavy  cast-iron  plate  bolted  to  a  concrete  foundation  of  sufficient 
mass  to  insure  stability. 

Fig.  117  gives  the  details  of  one  of  the  steel  chimneys  at  the  power 
house  of  the  South  Side  Elevated  Railroad,  Chicago,  111. 

133.  Thickness  of  Plates.  —  The  sheet  is  thickest  at  the  bottom, 
decreasing  toward  the  top  of  the  stack.  The  proper  thickness  for  any 
given  section  may  be  determined  by  treating  the  shaft  as  a  uniformly 
loaded  cantilever,  the  stresses  being  expressed  by  the  equation 

(42) 
in  which 

P  =  the  total  wind  pressure  in  pounds. 

h  =  length  of  the  chimney  in  inches  to  the  center  of  wind  pressure 
(h=  L/2  for  a  cylindrical  chimney). 

S  =  safe  stress.  A  low  value  of  6000  pounds  per  square  inch  for 
single-riveted  joints  and  8000  for  double-riveted  joints  is 
recommended,  for  the  reason  that  a  tube  of  such  large 
diameter  with  thin  walls  will  hardly  fail  by  rupture  according 
to  the  formula,  but  by  flattening  and  bending. 

-  =  sectional  modulus. 
e 

Dl  =  external  diameter  of  the  shell,  inches. 
D2  =  internal  diameter  of  the  shell,  inches. 

For  chimneys  under  7  feet  in  diameter  and  150  feet  in  height  the 
thickness  of  plate  should  not  be  less  than  -fg  inch,  nor  less  than  J  inch 
for  larger  sizes. 


CHIMNEYS 


221 


FIG.  117.  Steel  Chimney  at  the  Power  Hovise  of  the  South  Side  Elevated  Railroad,  Chicago, 


222 


STEAM  POWER  PLANT  ENGINEERING 


It  is  customary  to  make  the  courses  about  5  feet  in  height  for  con- 
venience in  erection. 

Table  29  gives  the  dimensions  of  self-supporting  steel  stacks  as  made 
by  the  Riter  Conley  Company  of  Pittsburg,  who  use  the  following 
empirical  formula  in  determining  the  thickness  of  the  shell 

(43) 

in  which 

Si  =  stress  per  lineal  inch  of  section  considered, 
M  =  wind  moment  in  inch-pounds,  and 
D^  =  diameter  of  the  shaft  in  inches. 

Allowing  8000  pounds  per  square  inch  as  the  safe  stress  for  single- 
riveted  joints  and  10,000  for  double-riveted  joints,  the  required  thick- 
ness is  found  by  dividing  /St  by  8000  or  10,000. 

Example :  Determine  the  thickness  of  plate  at  a  section  150  feet 
from  the  top  of  a  cylindrical  steel  stack  12  feet  in  diameter  and  200 
feet  high.  Horizontal  seams  to  be  double  riveted. 

The  total  wind  pressure  on  the  section  is 

150  X  12  X  25  =  45,000  pounds. 
The  moment  arm  is 
ijp-  X  12  =  900  inches. 

Di  =  144  inches;  S  =  8000  pounds  per  square  inch. 


TABLE   29. 
STEEL  STACKS.  —  SIZES  OF  RITER  CONLEY  COMPANY,   PITTSBURG. 


Diameter 
of  Flue. 

Total 
Height. 

Total 
Weight. 

How  Made. 

Ft.     In. 
5     6 

Ft. 
165 

Lb. 
67,000 

40  ft.  of  &  in.,  45  ft.  of  J  in.,  50  ft.  of  &  in.,  30  ft.  of 

fin. 

7  .0 

160 

79,000 

30  ft.  of  &  in.,  50  ft.  of  1  in.,  50  ft.  of  &  in.,  30  ft.  of 

fin. 

8     6 

150 

94,000 

60  ft.  of  i  in.,  60  ft.  of  &  in.,  30  ft.  of  f  in. 

10    0 

200 

150,000 

90  ft.  of  1  in.,  60  ft.  of  &  in.,  50  ft.  of  |  in. 

12    0 

200 

175,000 

35  ft.  of  i  in.,  35  ft.  of  &  in.,  35  ft.  of  &  in.,  35  ft. 

of  H  in.,  35  ft.  of  f  in.,  25  ft.  of  £f  in. 

11     6 

225 

232,000 

40  ft.  of  i  in.,  40  ft.  of  &  in.,  40  ft.  of  &  in.,  40  ft.  of 

H  in.,  40  ft.  of  f  in.,  25  ft.  of  &  in. 

12     0 

255 

256,000 

75  ft.  of  i  in.,  65  ft.  of  ^  in.,  55  ft.  of  f  in.,  35  ft. 

of  j$  in.,  25  ft.  of  \  in. 

CHIMNEYS  223 

Substituting  these  values  in  equation  (42), 


32V        D 
45,000  X  1800  =  8QOO  x   3.14  f 1444  -  Z)^ 


Da   =  143.36. 


Now 


2 
=  144  -  143.36 

2 
=  0.32  inch. 

The  nearest  commercial  size  lies  between  nine  thirty-seconds  and  five- 
sixteenths. 

The  Riter  Conley  formula  gives  for  this  section 

s    •        M       _  45,000  X  900 
1      0.8  D*         0.8  X  1442 
=  2440  pounds. 


134.  Riveting.  —  The  diameter  of  rivets  should  always  be  greater 
than  the  thickness  of  the  plate  but  never  less  than  one-half  inch.     The 
pitch  should  be  approximately  2£  times  the  diameter  of  the  rivet,  and 
always  less  than  16  times  the  thickness  of  the  plate.     Single-riveted 
joints  are  ordinarily  used  on  all  sections  except  the  base,  where  the 
joint  should  be  double  riveted  with  rivets  staggered,  although  in  very 
large  stacks  all  horizontal  seams  are  double  riveted  to  give  greater 
stiffness  to  the  shaft. 

135.  Stability  of  Steel  Chimneys.  —  The  wind  being  ordinarily  the 
only  force  tending  to  overturn  the  stack,  and  the  chimney  being  rigidly 
bolted  to  the  foundation,  a  condition  of  stability  requires  that 

(Wc  +  WF)  -  be  equal  to  or  greater  than  P  (—  +  h\  (44) 

\  2          / 

in  which 

Wc  =  weight  of  the  chimney  in  pounds. 

WF  —  weight  of  the  foundation. 

P  =  total  wind  pressure  in  pounds. 

D,  H,  and  h,  in  feet,  as  indicated  in  the  figure. 


224 


STEAM  POWER  PLANT  ENGINEERING 


Expressed  graphically:  Lay  off  GP,  Fig.  118,  equal  to  the  total  wind 
pressure  in  direction  and  amount  and  acting  at  the  center  of  pressure 

of  the  shaft;  lay  off  GW  equal  to  the 
weight  of  the  stack  and  foundation; 
find  the  resultant  GR  and  produce  it 
to  intersect  the  base  line  as  at  R';  if  R' 
falls  within  the  inner  third  of  the 
base  the  stack  is  stable,  provided,  of 
course,  that  the  chimney  is  properly 
designed  and  constructed.  Therefore 
the  heavier  the  combined  weight  of  the 
chimney  and  its  foundation  the  more 
stable  the  structure.  (See  also  para- 
graph 140.) 

D  in  Fig.  118  varies  from  one-tenth 
to  one-fifteenth  H,  depending  upon  the 
character  of  the  subsoil.  For  the 
ordinary  concrete  foundation,  Christie 
("Chimney  Design  and  Theory,"  p.  57) 
gives  as  an  average  value  for  D 


D  = 


26,000 


+  10. 


(46) 


FIG.  118. 


Detailed  Description  of  Steel  Chimneys :  Eng. 
Rec.,  Feb.  15,  1902,  p.  146,  July  5,  1902,  p.  2, 
April  23,  1904,  p.  523,  Sept.  10,  1904,  p.  314; 

Power,  Dec.,  1905,  p.  716,  April,  1905,  p.  231,  Jan.,  1902,  p.  6;  Engr.  U.S.,  Sept.  1, 
1904,  p.  591,  June  15,  1905,  p.  403;   Eng.  News,  July  20,  1905,  p.  64. 

136*  Brick  Chimneys.  —  By  far  the  greater  number  of  power-plant 
chimneys  are  of  brick  construction  and  usually  of  circular  section, 
though  octagonal,  hexagonal,  and  square  sections  are  quite  common. 
The  round  chimney  requires  the  least  weight  for  stability,  and  the  others 
in  the  order  mentioned.  Taking  the  total  wind  pressure  on  the  flat 
surface  of  a  square  stack  as  unity,  the  effective  pressure  for  the  same 
projected  area  will  be  0.75  for  the  hexagonal,  0.6  for  the  octagonal,  and 
0.5  for  the  round. 

Brick  chimneys  may  be  divided  into  two  general  classes : 

1.  Single  shell,  Fig.  119,  and 

2.  Double  shell,  Fig.  120. 

The  double  shell  is  the  most  common  and  consists  of  an  outer  shaft 
of  brickwork  and  an  inner  core  or  lining  extending  part  way  or 
throughout  the  entire  length  of  the  shaft. 


CHIMNEYS 


225 


/f'-S* 


/6-s» 


**'- 


TOP  OF 
FOUNDATION, 


30*0* 


TOTAL   HEIGHT 
ABOVE  FOUNDATION 

aoorr. 


8 


10 


20ft 


-y/v^ 

— H 


QECT/Oti  ON  A-A  SECTION  ON  B~B 

FIG.  119.  Custodis  Radial  Brick  Chimney. 


226 


STEAM  POWER  PLANT  ENGINEERING 


The  single  shell  is  the  general  construction  where  carefully  burned 
and  selected  brick  not  easily  affected  by  the  heat  are  used.  As  the 

inner  core  or  lining  is  independent 
of  the  outer  shell  and  has  no  part 
in  the  strength  of  the  chimney,  the 
rules  for  determining  the  thickness 
of  the  walls  are  practically  the  same 
for  both  single  and  double  shell. 

137.   Thickness  of  Walls.  —  The 
thickness  of  the  wall  should  be  such 
as  to  require  minimum  weight  of 
FIG.  H9a.  Custodis  Radial  Brick.          material  for  the  proper  degree  of 

stability,   due   consideration   being 

paid  to  the  practical  requirements  of  construction.  The  thickness 
does  not  vary  uniformly,  but  decreases  from  bottom  to  top  by  a 
series  of  steps  or  courses  as  in  Fig.  121.  In  general,  the  thickness  at 
any  section  should  be  such  that  the  resultant  stress  of  wind  and  weight 
of  shaft  will  not  put  the  masonry  in  tension  on  the  windward  side  or 
in  excessive  compression  on  the  leeward  side. 

For  circular  chimneys  using  common  red  brick  for  the  outer  shell 
the  following  approximate  method  gives  results  in  conformity  with 
average  practice: 

t  =  4  +  0.05  d  +  0.0005  #,  (47) 

where 

t  =  thickness  in  inches  of  the  upper  course,  neglecting  ornamenta- 
tion, and  should,  of  course,  be  made  equal  to  the  nearest 
dimension  of  the  brick  in  use.  Ordinary  red  bricks  measure 


d 


clear  inside  diameter  at  the  top,  inches. 
height  of  stack,  inches. 


Beginning  at  the  top  with  this  thickness,  add  one-half  brick,  or 
4  inches,  for  each  25  or  30  feet  from  the  top  downwards,  using  a  batter 
of  1  in  30  to  1  in  36. 

The  minimum  value  of  t  for  stacks  built  with  inside  scaffolding 
should  be  7  inches  for  radial  brick  and  8£  inches  for  common  brick, 
as  a  thinner  wall  will  not  support  the  scaffold.  Radial  brick  for  chim- 
neys are  made  in  several  sizes,  so  that  the  thickness  of  the  walls  when 
they  are  used  increases  by  about  2  inches  at  the  offsets. 

For  specially  molded  radial  brick  or  for  circular  shells  reenforced 
as  in  Fig.  120  the  length  of  the  different  courses  may  be  much  less  than 


CHIMNEYS 


227 


FIG.  120.   Brick  Chimney  at  the  Power  Plant  of  the  Armour  Institute  of  Technology. 


228 


STEAM  POWER  PLANT  ENGINEERING 


stated  above.  The  external  form  of  the  top  is  a  matter  of  appearance, 
and  may  be  designed  to  suit  the  taste,  but  should  be  protected  by  a  cast- 
iron  or  tile  cap  and  provided  with  lightning  rods.  Ladders  for  reach- 
ing the  top  of  the  chimney  are  generally  located  inside  of  brick  stacks 
and  outside  of  steel  ones. 

Professor  Lang's  rule  (Eng.  Rec.,  July  20,  1901,  p.  53)  for  determining 
the  length  of  the  different  courses  is  (Fig.  121) 


h  = 


20  t  +60  *  +  0.1056G  +  2.5^-  +  656  tan  a  -  0.007  H 

2 

-  0.453  p  -  18 


in  which 


(48) 


h  =  length  of  the  course  under  consideration. 
C  =  constant  =  1  for  a  circular,   0.97  for  an 

octagonal,  and  0.83  for  a  square  chimney. 
i  =  increase  in  thickness  for  each  succeeding 

section  in  feet. 

(7  =  weight  per  cubic  foot  of  brickwork. 
p  =  wind  pressure,  pounds  per  square  foot. 
a  =  angle  of  the  internal  batter. 
All  other  notations  as  indicated  in  Fig.  121. 

For  chimneys  over  100  feet  in  height  he 
recommends  that  100  be  used  instead  of  the 
actual  height,  since  the  critical  point  will  be  in 
one  of  the  lower  sections  and  not  at  the  base. 

If  a  value  of  h  is  obtained  which  is  not  con- 
tained an  even  number  of  times  in  H,  it  may  be 
slightly  increased  or  decreased  so  as  to  effect 
this  result. 

To  determine  the  stresses  at  any  section  the 
shaft  is  treated  as  a  cantilever  uniformly  loaded 
with  a  maximum  wind  pressure  of  25  pounds  per 
square  foot.  If  the  tension  on  the  windward 

side  subtracted  from  the  compression  leaves  a  positive  remainder,  the 
chimney  will  be  stable;  if  the  remainder  is  negative,  the  masonry  will 
be  in  tension,  which  it  withstands  but  feebly.  The  sum  of  the  com- 
pressive  stresses  on  the  leeward  side  due  to  wind  pressure  and  weight 
must  be  less  than  the  crushing  strength  of  the  masonry.  The  practice, 
however,  of  assuming  a  fixed  value  for  allowable  pressure  irrespective 
of  the  height  of  the  stack  gives  dimensions  that  are  too  low  for  small 


CHIMNEYS  229 

stacks  and  too  high  for  large  stacks.  According  to  Professor  Lang, 
compressive  stress  on  the  leeward  side  in  pounds  per  square  inch  with 
single  chimneys  should  not  exceed 

p  =  71  +  0.65  L,  (49) 

where 

p  =  pressure  in  pounds  per  square  inch. 

L  =  distance  in  feet  from  top  of  chimney  to  the  section  in  question. 

With  double  shell  p  =  85  +  0.65  L.  (50) 

The  tension  on  the  windward  side  should  not  exceed, 

For  single  shell:  p  =  (18.5  +  0.056  L).  (51) 

For  double  shell:  p  =  (21.3  +  0.056  L).  (52) 

Example:  Determine  the  maximum  stress  in  the  outer  fibers  of 
the  brickwork  at  the  base  of  section  8  of  the  chimney  illustrated  in 
Fig.  119  when  the  wind  is  blowing  100  miles  an  hour.  Assume  the 
weight  of  the  brickwork  120  pounds  per  cubic  foot. 

A  wind  velocity  of  100  miles  per  hour  is  estimated  to  exert  a  pressure 
of  50  pounds  per  square  foot  on  a  flat  surface  and  approximately  25 
pounds  per  square  foot  of  projected  area  on  a  cylindrical  surface.  The 
height  of  the  chimney  to  section  8  is  131.4  feet.  The  projected  area 
as  computed  from  the  figure  is  1800  square  feet.  Hence  p,  the  total 
wind  pressure,  is  1800  X  25  =  45,000  pounds.  The  volume  of  brick- 
work above  section  9  may  be  calculated,  and  is  6150  cubic  feet,  hence 
the  weight  W  =  6150  X  120  =  738,000  pounds. 

The  area  of  the  joint  at  this  section  is  75.3  square  feet,  therefore  the 
pressure  due  to  the  weight  of  the  superimposed  brickwork  is  738,000 
divided  by  75.3  =  9800  pounds  per  square  foot.  To  find  the  stress 
due  to  the  wind  pressure,  substitute  the  proper  values  in  equation  (42) : 

Ph  =  S  -  =0.0983  (^±— - 
e  \      Z>x 

Here 

P  =  45,000  as  computed  above. 

h  =  55  feet.      (Found  by  laying  out  the  section  and  locating  the 

center  of  gravity.) 
D1==  16.2. 
D  =  12.9. 

Whence 

1  a  94         1  o  Q4 

45,000  X  55  =  0.0983  ib^   ~  LZ'*  8. 
From  which  S  =  9907  pounds  per  square  foot. 


230  STEAM  POWER  PLANT  ENGINEERING 

The  net  stress  on  any  part  of  the  section  is  the  resultant  of  that  due 
to  the  weight  of  the  stack  and  that  caused  by  the  wind,  the  net  stress 
on  the  windward  side  being 

9907-  9800  =  107  pounds  per  square  foot, 

which  is  evidently  a  tensile  stress  and  should  never  exceed  the  value 
given  by  formula  (51): 

p  =  (18.5  +  0.056  L) 

=  (18.5  +  0.056  X  131.4) 

=  25.8  pounds  per  square  inch 

=  3715  pounds  per  square  foot. 

The  net  compressive  stress  on  the  leeward  side  is  9800  +  9907 
=  19,707  pounds  per  square  foot,  which  should  not  exceed  that  given 
by  formula  (49) : 

p  =  71  +  0.65  L 

=  71  +  0.65  X  131.4 

=  156.4  pounds  per  square  inch 

=  22,521  pounds  per  square  foot. 

(See  also  analysis  of  steel-concrete  chimney,  paragraph  142.) 

138.  Core  and  Lining.  —  The  core  or  lining  of  a  brick  chimney  is 
commonly  carried  to  the  top  of  the  shaft,  though  it  sometimes  extends 
only  part  of  the  distance.     The  inside  diameter  is  generally  uniform, 
the  offsets   being  made  on  the  outside.      The   core  and  outer  shell 
should  be  independent  to  prevent  injury  due  to  expansion  of  the  core. 
The  rules  for  the  thickness  of  lining  in  steel  chimneys  apply  also  to  brick 
chimneys.     The  batters  for  the  inner  and  outer  shells  should  be  such 
as  to  allow  at  least  2  inches  clearance  between  the  two  shafts  at  the 
top,  and  the  top  should  be  protected  by  an  iron  ring  or  by  a  projecting 
ledge  from  the  outer  shell. 

139.  Materials  for  Brick  Chimneys.  —  Brick  for  the  external  shaft 
should  be  hard  burned,  of  high  specific  gravity,  and  laid  with  lime 
mortar  strengthened  with  cement.     Lime  mortar  itself  is  more  resist- 
ant to  heat,  but  hardens  slowly  and  may  cause  distortion  in  newly 
erected  stacks,  and  hence   should  be  used  only  when   a  long  time  is 
taken  in  building.     Mortar  of  cement  and  sand  alone  is  not  to  be 
recommended,  since  it  does  not  resist  heat   well  and  is  attacked  by 
carbon  dioxide,  particularly  in  the  presence  of  moisture.      A  mortar 
consisting  of  1  part  by  volume  of  cement,  2  of  lime,  and  6  of  sand 


CHIMNEYS 


231 


may  be  used  for  the  upper  brickwork,  1,  2£,  and  8  respectively  for  the 
lower  part,  and  1,1,  and  4  respectively  for  the  cap.  The  harder  the 
brick  the  more  cement  is  necessary,  as  lime  does  not  cling  so  well  to 
hard,  smooth  surfaces.  The  inner  core  may  be 
constructed  of  second-class  fire  brick,  since  the 
temperature  seldom  exceeds  600  degrees  F. 
Lime  mortar  is  invariably  used  for  the  core. 
140.  Stability  of  Brick  Chimneys.  —When  there 
is  no  wind  blowing  and  the  chimney  is  built  sym- 
metrically about  a  vertical  axis  the  pressure 
due  to  weight  is  uniformly  distributed  over  the 
x  bearing  surfaces,  and  the  center 

of  pressure  lies  in  the  line  XX, 
Fig.  122.  But  when  the  wind 
blows  the  pressure  exerted  tends 
to  tilt  the  shaft  as  a  whole- 
column  in  the  direction  of  the 
current,  and  the  pressure  de- 
creases from  the  windward  to 
the  leeward  side  of  the  base, 
until,  with  a  sufficiently  high 
velocity  of  wind,  it  may  become 
zero,  in  which  case  the  center 
of  pressure  moves  a  distance  q 
towards  the  leeward  side  of 
the  base.  As  soon  as  the  pressure  at  A  becomes  zero  the  joint  begins 
to  open  (assuming  no  adhesion  between  chimney  and  base)  and  the 
shaft  is  evidently  in  the  condition  of  least  stability.  The  distance  q 
through  which  the  center  of  pressure  has  moved  is  called  the  radius 
of  the  statical  moment.  For  any  column  it  may  be  shown  that 

q  =  JL  .     (Rankine,  "  Applied  Mechanics,"  p.  229.)  (53) 

Ae 

in  which 

I  =  moment  of  inertia  of  the  section. 
A  =  area  of  the  section. 

e  =  distance  from  the  center  of  the  shaft  to  the  outer  edge  of  the 
joint. 

D 


122. 


Thus  for  circular  section, 
For  square  section, 


8"T 


232  STEAM  POWER  PLANT  ENGINEERING 

For  annular  circular  ring,  q  =  — . 

8  D 

For  hollow  square,  q  =    . 

QD 

The  relationship  between  weight  of  shaft  and  wind  pressure  for  the 
condition  of  least  stability  is 

Ph  =  Wq,  (54) 

in  which 

P  =  total  wind  pressure,  pounds. 

h  =  distance  in  feet  from  the  base  line  of  the  section  under  con- 
sideration to  center  of  gravity  of  that  section. 
W  =  weight  of  shaft  in  pounds  above  the  assumed  base  line. 
q  =  radius  of  the  statical  moment. 

The  condition  of  least  stability  for  round  chimneys  requires,  there- 
fore, that 

Ph  =  W  J2!_±_^l  (55) 

For  many  purposes  it  is  sufficiently  accurate  to  assume  D  =  d,  and 
equation  (55)  becomes 

Ph  =  W  —  for  round  chimneys.  (56) 

Ph  =  W — for  square  chimneys.  (57) 

3 

The  rule  commonly  used  in  Germany,  and  which  is  finding  much 
favor  with  engineers  in  the  United  States,  gives  for  the  condition  of 
least  stability 

R  +  7r )  =  ph-   (En9-  Rec->  July  27>  1901>  P-  82-)       (58) 


Notations  as  in  Fig.  122,  all  dimensions  in  feet. 

This  permits  of  a  lighter  chimney  than  equation  (55),  and  the  maxi- 
mum wind  pressure  may  be  assumed  to  put  the  joint  on  the  wind- 
ward side  in  tension  or  even  to  permit  a  slight  opening  of  same. 

A  rule  of  thumb  for  stability  is  to  make  the  diameter  of  the  base  one- 
tenth  of  the  height  for  a  round  chimney;  for  any  other  shape  to  make 
the  diameter  of  the  inscribed  circle  of  the  base  one-tenth  of  the  height. 

The  factor  of  stability  is  the  quotient  obtained  by  dividing  the  value 
of  q  from  formula  (54)  by  that  from  (53).  If  less  than  unity,  the 
chimney  is  in  tension  at  the  outer  fiber  on  the  windward  side,  and  must 
be  redesigned  unless  the  tension  is  less  than  that  allowed  by  equation  (51). 
Calculations  for  stability  should  be  made  for  various  sections. 


CHIMNEYS  233 

Example:  Analyze  the  chimney  illustrated  in  Fig.  119  for  stability 
at  say  section  8,  the  following  data  referring  to  the  portion  above  the 
base  line  of  this  section. 

From  the  drawing  : 

Projected  area  of  the  stack,  1800  square  feet. 

Volume  of  brickwork,  6150  cubic  feet. 

Outside  diameter  of  base,  16.2  feet. 

Inside  diameter  of  base,  12.9  feet. 

Center  of  pressure  to  base  line,  55  feet. 

Total  height  above  base  line,  131.4  feet. 

Maximum  total  wind  pressure  : 

P  =  1800  X  25  =  45,000  pounds. 

Weight  of  shaft  : 

W  =  6150  X  120  =  738,000  pounds. 

For  stability,  according  to  equation  (55), 

' 


Substituting  the  proper  values  : 

Ph  =  45,000  X  55  =  2,475,000  foot-pounds. 

=  2>441'000' 


While  Ph  is  slightly  greater  than  W  -  —  —  ,  for  practical  purposes 

8  D 

the  shaft  at  this  section  would  be  called  stable  under  maximum  allow- 
able wind  pressure. 

For  stability,  according  to  equation  (58), 

Ph  <  W  (%R  +  Jr). 

Ph  =  2,475,000,  as  determined  above. 

R  +  i  r)  =  738,000  @     + 


=  4,177,000. 

Ph  is  therefore  considerably  less  than  W(%  R  +  Jr),  and  the  con- 
dition imposed  in  equation  (58)  is  more  than  fulfilled. 

Detailed  Description  of  Brick  Chimneys  :  At  Paris  Exposition,  Eng.  Rec.,  Feb. 
17,  1900,  p.  155;  Boston  Elevated  Ry.,  Eng.  Rec.,  Dec.  22,  1900,  p.  593;  Plymouth 
Cordage  Co.,  Eng.  Rec.,  May  18,  1901,  p.  466;  Boston,  Edison  Co.,  Eng.  Rec.,  Oct. 
10,  1903,  p.  438;  Smelting  Works,  Freiburg,  Germany,  Power,  Aug.,  1900,  p.  13;  Ter- 
minal R.R.  Assn.  of  St.  Louis,  Mo.,  Cassier's  Mag.,  Jan.,  1898,  p.  261;  Cambridge 
Electric  Light  Co.,  Engr.  U.S.,  May  15,  1904,  p.  331;  Interborough  Rapid  Transit 
Co.,  Engr.  U.S.,  Nov.  1,  1904,  p.  737;  Metropolitan  St.  Ry.  Co.,  Power,  March,  1899, 
p.l. 


234  STEAM  POWER  PLANT  ENGINEERING 

141.  Custodis  Radial  Brick  Chimney.  —  Fig.  119  gives  the  details  of 
a  200  x  10  foot  radial  brick   chimney  constructed  of  special    molded 
radial  brick,  formed  to  suit  the  circular  and  radial  lines  of  each  section, 
thus  permitting  them  to  be  laid  with  thin,  even  mortar  joints.     The 
blocks  are  much  larger  than  common  brick  and  the  number  of  joints  is 
proportionately  reduced.     They  are  molded  with  vertical  perforations, 
as  shown  in  Fig.  119a,  which  permits  thorough  burning,  thereby  increas- 
ing the  density  and  strength  and  at  the  same  time  reducing  the  weight 
of  the  block.     In  laying,  the  mortar  is  worked  into  the  perforations 
about  one-half  inch.     The  first  60  feet  above  the  base  are  octagonal  in 
section,  with  36-inch  walls,  and  the  balance  of  circular  section,  with  walls 
tapering  gradually  from  22  inches  to  7}  inches  in  thickness.     A  radial 
brick  lining  extends  60  feet  from  the  base  as  indicated.     The  chimney 
was  designed  to  furnish  draft  for  a  3500-horse-power  boiler  plant  and 
cost,  erected,  $8,800.     The  entire  weight  of  the  chimney  exclusive  of 
foundation  is  870  tons. 

Radial  brick  chimneys  without  the  inner  lining  are  likely  to  be 
unduly  affected  by  heat. 

The  tallest  chimney  in  the  world  (1907),  located  at  Great  Falls, 
Mont.,  is  of  the  Custodis  type,  and  is  used  for  leading  off  the  gases  from 
the  smelter  plant  of  the  Boston  and  Montana  Consolidated  Copper  and 
Silver  Mining  Company.  The  height  above  the  top  of  the  foundation  is 
506  feet,  and  the  internal  diameter  at  the  top  50  feet.  The  chimney 
and  foundation  cost  approximately  $200,000. 

Custodis  Chimney  Details:  Eng.  Rec.,  Oct.  1,  1904,  p.  385;  Power,  May,  1900, 
p.  12. 

142.  Steel-Concrete  Chimneys.  —  The  use  of  concrete  reenforced  with 
iron  or  steel  for  the  construction  of  chimneys  is  rapidly  increasing.    The 
advantages  claimed  for  this  class  of  stack  are: 

1.  Light  weight  of  the  whole  structure,  being  but  one-third  as  great 
as  an  equivalent  common  brick  chimney.     The  space  occupied  is  much 
less  than  with  either  brick  or  steel  stack,  on  account  of  the  thinness  of 
walls  at  the  base  and  the  absence  of  any  flare  or  bell. 

2.  Total  absence  of  joints,  the  entire  structure  including  foundation 
being  a  monolith. 

3.  Great  resisting  power  against  tension  and  compression. 

4.  Rapidity  of  construction.     May  be  erected  at  an  average  rate  of 
six  feet  per  day. 

5.  Adaptability  of  the  material  to  any  form. 

This  type  of  chimney  being  comparatively  new,  little  data  concern- 
ing depreciation  are  available,  but  some  which  have  been  in  use  ten 
years  show  little  or  no  deterioration. 


CHIMNEYS 


FIG.  123.  Weber  Reenforced  Concrete  Chimney. 


236 


STEAM  POWER  PLANT  ENGINEERING 


130 


158.0 


i 


Fig.  123  gives  the  details  of  a  Weber  steel-concrete  chimney  erected 
at  Portland,  Oregon,  for  the  Portland  General  Electric  Company.  The 
entire  structure,  foundation  and  shaft,  is  a  monolith,  238  feet  in  total 
height  and  12  feet  internal  diameter,  weighing  only  889  tons.  It 
occupies  but  168  square  feet  of  ground  space  at  the  grade  level.  The 
weight  not  including  foundation  is  470  tons.  The  stack  was  erected 
complete  in  58  working  days,  and  cost  approximately  $13,000. 

The  cement  used  was  German  Portland  mixed  with  select  bank  sand 
in  proportion  of  one  to  three,  gravel  or  crushed  stone  being  used  only 
in  the  foundation  below  the  ground.  The  mortar  was  used  medium 
dry  and  tamped  in  the  form  around  the  steel  reinforcement. 

The  shaft  is  of  the  double-shell  type,  with  inner  core  extending  70 
feet  above  the  grade.  The  core  is  but  4  inches  in  thickness  at  the 

base,  and  the  outer  shell  8  inches. 
Both  inner  and  outer  shell  are 
reenforced  with  vertical  T  bars, 
IJxl-Jx-^  inch,  of  low-carbon 
Bessemer  steel,  spaced  at  the  base 
24  inches  between  centers  in  the 
inner  core  and  4  inches  in  the 
outer  shell,  and  increasing  in 
spacing  to  the  top,  where  the  dis- 
tance between  the  bars  is  12 
19.75  sq.rt.  inches.  The  horizontal  rings  are 

238.0  ^  ;  1  x  1  x  J    T's    spaced    18    inches 

between  centers  in  the  core  and 
36  inches  in  the  outer  shell.  The 
steel  bars  vary  from  16  to  30  feet 
in  length,  and  where  they  meet 
lengthwise  are  lapped  not  less 
than  24  inches.  The  use  of  differ- 
ent lengths  of  steel  prevents  the 
laps  from  concentrating  in  any 
given  section. 

The  tallest  chimney  of  this  type 
(1907)  was  erected  for  the  Butte 
Reduction  Works  at  Butte,  Mont. 
Its  height  is  350  feet  and   inside 
FIG.  124.  diameter  18  feet. 

The  following  strain  sheet  gives  the  Weber  Company's  analysis  of 
the  chimney  illustrated  in  Fig.  123,  and  is  based  on  a  wind  pressure 
of  50  pounds  per  square  foot.  Notations  as  in  Fig.  124. 


CHIMNEYS  237 

WEIGHTS. 

=  weight  of  foundation 

150 

2—  3)150 

=  523,200  pounds. 

150  =  weight  per  cubic  foot  of  concrete. 
We  =  earth  weight  on  foundation 

=  {lihQ  —    (volume  of  foundation)}  100 

=  (7200  cubic  feet  -  3995  cubic  feet)  100 

=  320,500  pounds. 

100  =  weight  per  cubic  foot  of  earth. 
W  =  weight  of  shaft 

=  {At(/i4  +  hs)  +  A2  (hi  +  &3)  +  Aji^  150 

=  {38.5  (72  +3)  +  13  (72  +  3)  +  19.75-158}  150 

=  934,950  pounds. 
Wt  =  total  weight 

=  Wf  +  We  +  W  =  1,778,650  pounds  (889  tons). 

SECTION  AT  GRADE  Gr. 

I.       W%  =  weight  of  outer  shell  and  single  shell  above  section 
=  (AJii  +  AJi^)  150 
=  (28.5  •  72  +  19.75  •  158)  150 
=  775,806  pounds. 

II.          r  =  radius  of  statical  moment 


A 

14.66 

8 
=  3.35  feet. 

III.        P  =  wind  pressure  on  chimney 
D  h  ^-  +  D  h  — 

=  14.66  X  72  X  25  + 13  X  158  X  25 
=  77,738  pounds. 


238  STEAM  POWER  PLANT  ENGINEERING 

M  =  wind  moment  on  section 


=  14.66  X  72  X  y-  X  —-  +(l3  X  158  X  |°-)(72  + 
=  8,703,818  foot-pounds. 


IV.        N  =  statical  moment 
=  rWi. 

=  3.35  X  775,806 
=  2,598,950  foot-pounds. 

V.         B  =  bending  moment 
-  M  -N 

=  8,703,818  -  2,598,950 
=  6,104,868  foot-pounds. 

VI.          —  =  section  modulus 


•  Q.0982  X 


=  169,703. 
VII.          z  =  tension  per  square  inch  sectional  area 

=  125-^- 
e 

=  12  X  6,104,868  -^  169,703 
=  432.5  pounds. 

VIII.          Z  =  total  tension 
=  144  A& 

=  144  X  28.5  X  432.5 
=  1,825,015  pounds. 

IX.          s  =  area  steel  required 

ry 

=  _  .  (a  =  sectional  strain  on  steel) 

a 

=  16,000  pounds  per  square  inch 

=  114.2  square  inches. 


CHIMNEYS  239 

X.      K  =  number  of  bars 


L  .   (#  =  0.45  square  inches  =  area  of  one  bar) 


=  252  bars. 

FOR  STABILITY. 

XI.        L  =  length  of  one  side  of  base. 


=  8,703,818       6 

1,778,650 
=  29.4  feet. 

SECTION  42'  0"  ABOVE  GRADE. 
I.     Wl  =  596,250  pounds. 
II.      r    =  3.35  feet. 

III.  P    =  62,295  pounds.     M=  5,761,325  foot-pounds. 

IV.  N    =  1,997,438  foot-pounds. 
V.     B     =  3,763,889  foot-pounds. 

VI.     _      =  169,703. 

e 

VII.  z  =  222  pounds. 

VIII.  Z  =  91  1,088  pounds. 

IX.  s  =57  square  inches. 

X.  K  =  127  bars. 

SECTION  AT  OFFSET. 

I.     Wl   =  468,000  pounds. 
II.      r     =  3  feet. 

III.  P    =  51,350  pounds.     M=  4,056,650  foot-pounds. 

IV.  N    =  1,404,000  foot-pounds. 
V.     B     =  2,652,650  foot-pounds. 

VI.     {      =  102,041. 

e 

VII.  z  =  311  pounds. 

VIII.  Z  =  786,000  pounds. 

IX.  s  =55.5  square  inches. 

X.  K  =  123  bars. 


240  STEAM   POWER  PLANT  ENGINEERING 

SECTION  50'  0"  FROM  TOP. 

I.     Wl  =  148,125  pounds. 
II.       r    =  3  feet. 

III.  P     =  16,250  pounds.     M=  406,250  foot-pounds. 

IV.  N    =  444,365  foot-pounds. 

Since  the  statical  moment  N  is  greater  than  the  wind  moment  M, 
there  is  no  bending  moment  B,  so  no  steel  is  required,  the  chimney 
above  this  section  standing  of  its  own  weight.  However,  thirty-two 
bars  are  continued  to  the  top. 

Detailed  Description  of  Concrete-Steel  Chimneys :  Weber  Chimney,  Power,  July, 
1907,  p.  476;  Burt  Portland  Cement  Co.,  Eng.  News,  Dec.  29,  1904,  p.  579;  Leiter 
Coal  Mines,  Eng.  Rec.,  March  21,  1904,  p.  661,  Power,  Dec.,  1904,  p.  787;  Pacific 
Electric  Ry.  Co.,  Los  Angeles,  Cal.,  Eng.  News,  April  2,  1903,  p.  308;  Laclede  Fire 
Brick  Co.,  Eng.  News,  April  2,  1903,  p.  310;  Tall  Concrete  Chimneys,  Eng.  News, 
July  20,  1905,  p.  57,  Aug.  3,  1905,  p.  120,  Feb.  15,  1906,  p.  165,  Oct.  11,  1906, 
p.  387;  Butte  Reduction  Works,  Eng.  Rec.,  Feb.  3,  1906,  p.  124;  New  Types,  Eng. 
Rec.,  Dec.  15,  1906,  p.  670. 

143.  Breeching.  —  The  area  of  the  flue  or  breeching  leading  from  the 
boilers  to  the  chimney  is  generally  made  equal  to  or  a  little  larger  than 
the  internal  area  of  the  chimney,  20  per  cent  greater  being  an  average 
figure.     The  flue  may  be  carried  over  the  boilers  or  back  of  the  setting, 
or  even  under  the  fire-room  floor,  but  in  any  case  should  be  as  short  as 
possible  and  free  from  abrupt  turns.      Short  right-angled  turns  reduce 
the  draft  approximately  0.05  inch  for  each  turn,  and  a  convenient  rule 
is  to  allow  0.1  inch  loss  for  each  100  feet  of   flue  if  of   circular  cross 
section  and  constructed  of  steel,  and  double  this  amount  for  brick  flues 
of  square  section.     The  cross  section  of  the  flue  need  not  be  the  same 
throughout  its  entire  length,  but  may  be  tapered  and  proportioned  to 
the  number  of  boilers.     Where  two  flues  enter  the  stack  on  opposite 
sides,  a  diaphragm  is  inserted  as  indicated  in  Fig.  119.     Flues  should 
be  covered  with  heat-insulating  material. 

144.  Chimney  Foundations.  —  On  account  of  the  concentration  of 
weight  on  a  small  area  the  foundation  of  a  chimney  should  be  carefully 
designed.     In  most  cities  the  building  laws  limit  the  maximum  loads 
allowed  for  various  soils  and  materials,  and  although  they  vary  con- 
siderably the  average  is  approximately  as  follows: 

MATERIAL.  SAFE  LOAD,  LB.  PER  SQ.  FT. 

Hard-burned  brick  masonry,  cement  mortar,  1  to  2 20,000-30,000 

Hard-burned  brick  masonry,  cement  mortar,  1  to  4 18,000-24,000 

Hard-burned  brick  masonry,  lime  mortar 10,000-16,000 

Concrete,  Ito8 8,000-10,000 


CHIMNEYS  241 

KIND  OF  SOIL.  SAFE  LOAD,  TONS  PER  SQ.  FT. 

Quicksands  and  marshy  soils    0.5 

Soft  wet  clay 1.0 

Clay  and  sand  15  feet  or  more  in  thickness    1.5 

Pure  clay  15  feet  or  more  in  thickness 2.0 

Pure  dry  sand  15  feet  or  more  in  thickness   2.0 

Firm  dry  loam  or  clay  3.0-  4.0 

Gravel  well  packed  and  confined     6.0-  8.0 

Rock  broken  but  well  compacted    10.0-15.0 

Solid  bed  rock     Up  to  £  of  its  ultimate  crushing  strength. 

Tons  per  Pile. 

Piles  in  made  ground   2.0 

Piles  driven  to  rock  or  hardpan   25.0 

Chimney  foundations  as  a  rule  are  constructed  of  concrete  except 
where  the  low  sustaining  nature  of  the  soil  necessitates  the  use  of  piles 
or  a  grillage  of  timber  or  steel.  For  masonry  chimneys  the  foundation 
is  designed  to  give  the  necessary  support  to  the  shaft  without  particular 
reference  to  its  mass  or  distribution,  as  the  shape  of  the  foundation  has 
virtually  no  effect  on  its  stability  as  a  column.  In  steel  and  reenforced 
concrete  chimneys  the  shape  and  weight  of  the  foundation  are  a  function 
of  the  desired  factor  of  stability,  since  the  shaft  is  securely  anchored 
to  the  foundation  and  the  two  form  practically  one  mass.  The  founda- 
tion should  be  designed  to  fulfill  the  conditions  in  formula  (46)  in 
addition  to  the  requirements  for  mere  support. 

Table  30  gives  the  least  diameter  and  depth  of  foundation  for  steel 
chimneys  of  various  diameters  and  heights. 

145.  Chimney  Efficiencies.  —  The  chimney  as  a  mover  of  air  has  a 
very  low  thermodynamic  efficiency.  Compared  with  that  of  a  fan  its 
performance  is  very  poor,  and  mechanical -draft  concerns  sometimes 
use  this  as  an  argument. 

Example:  A  chimney  200  feet  high  and  10  feet  in  diameter  furnishes 
draft  for  a  battery  of  boilers  rated  at  3500  horse  power.  Average 
outside  temperature  60  degrees  F. ;  temperature  of  flue  gases  500  degrees 
F.;  calorific  value  of  the  fuel  14,000  B.T.U.  per  pound.  Compare  the 
thermal  efficiency  of  the  chimney  as  a  mover  of  air  with  that  of  a  forced 
draft  apparatus  of  equivalent  capacity. 

From  Table  24  we  find  that  a  chimney  200  feet  high,  with  tempera- 
tures as  stated  above,  will  furnish  a  theoretical  draft  of  1.27  inches, 
equivalent  to  a  pressure  of  6.6  pounds  per  square  foot.  Neglecting 
friction  the  height  H  of  a  column  of  external  air  which  would  produce 
this  pressure  is 

(59) 


242  STEAM  POWER  PLANT  ENGINEERING 

in  which 

h    =  height  of  the  chimney  in  feet. 

d    =  density  of  the  hot  gases  in  the  stack. 

dl  =  density  of  the  outside  air. 

Substitute  in  (59), 

dl  =  0.0763,     d  =  0.0435,     and  h  =  200. 

TABLE  30. 

SIZES    OF    FOUNDATION    FOR    STEEL    CHIMNEYS. 


Diameter,  Feet. 

Height,  Feet. 

Least  Diameter  of 
Foundation. 

Least  Depth  of 
Foundation. 

3 

100 

15'     9*. 

6'     0' 

4 

100 

16'     4* 

6'     0* 

4 

125 

18'     5" 

7'     0" 

5 

150 

20'     4* 

9'     0" 

5 

200 

23'     8* 

10'     0" 

6 

150 

21'  10* 

8'     0* 

6 

200 

25f     0" 

10'     0* 

7 

150 

22'     7* 

9'     0* 

7 

250 

29'     8* 

12'     0* 

9 

150 

23'     8* 

10'     0' 

9 

275 

33'     6* 

12'    0* 

11 

250 

24'     8* 

10'     0' 

11 

350 

36'     0" 

14'     0* 

).Q763  -  O.Q435\  onn 
0.0763         720< 
=  85.9  feet. 

The  theoretical  velocity  of  the  air  entering  the  base  of  the  chimney 
under  this  head  is 

v  =  V2gh 

=  \/2  X  32.2  X  85.9 
=  74.5  feet  per  second. 

The  weight  of  the  gas  escaping  per  second 

=  74.5  X  area  of  the  stack  X  0.0763 
=  446  pounds. 

The  displacement  of  this  volume  of  gas  is  the  result  of  heating  it  from 
60  to  500  degrees  F.  Taking  the  specific  heat  of  the  gas  as  0.2375,  the 
heat  necessary  to  displace  456  pounds  per  second  is 

Heat  required  =  446  X  0.2375  X  (500  -  60) 
=  46,500  B.T.U.  per  second 


CHIMNEYS 


243 


The  work  actually  performed  is  that  of  overcoming  a  total  resistance 
of  6.6  X  78.5  =  518  pounds  (78.5  =  internal  area  of  the  chimney) 
through  a  space  of  74.5  feet;  i.e., 

Work  done  =  74.5  X  518  =  38,591  foot-pounds  per  second, 
=  49.7  B.T.U.  per  second. 

Efficiency  =    49'7    =  .00107,  or  about  TV  of  1  per  cent. 
4t),ouu 

If  a  fan  be  substituted  for  the  chimney  and  we  allow  say  8  per  cent 
for  the  efficiency  of  engine  and  boiler,  40  per  cent  for  the  fan,  and  25  per 
cent  for  friction,  the  combined  efficiency  will  be 

0.08  X  0.40  X  0.75  =  0.024,  or  2.4  per  cent. 


The  fan  then  will  be 


0.024 
0.00107 


=  22.4  times  more  efficient  than  the 


chimney  as  a  mover  of  air. 
146.   Cost  of  Chimneys.  —  Christie  ("Chimney  Design  and  Theory") 

gives  the  following  costs  of  chimneys  150  feet  high  and  8  feet  internal 

diameter : 

Common  red  brick approximate  cost    $8,500.00 

Radial  brick do  do       6,800.00 

Steel,  self-supporting,  full  lined  do  do        8,300.00 

Steel,  self-supporting,  half  lined do  do        7,800.00 

Steel,  self-supporting,  unlined do  do       5,820.00 

Steel,  guyed do  do        4,000.00 

The  following  approximate  costs  of  various  sizes  of  a  well-known 
radial  brick  chimney  give  an  idea  of  the  variation  in  cost  due  to  in- 
crease in  diameter  and  height  : 


Size  of  Chimney. 

fVkCt 

Size  of  Chimney. 

Pn«st 

Height. 

Diameter. 

OOST/. 

Height. 

Diameter. 

vA)ot» 

Feet. 

Feet. 

Feet. 

Feet. 

75 

4 

$1,350.00 

175 

8 

$7,050.00 

75 

6 

1,950.00 

175 

10 

7,925.00 

75 

8 

2,650.00 

175 

12 

8,950.00 

75 

10 

3,725.00 

175 

14 

9,725.00 

125 

6 

3,500.00 

200 

8 

9,250.00 

125 

8 

4,250.00 

200 

10 

10,500.00 

125 

10 

4,675.00 

200 

12 

11,100.00 

125 

12 

5,125.00 

200 

14 

12,500.00 

150 

8 

6,150.00 

250 

10 

16,500.00 

150 

10 

7,125.00 

250 

12 

18,250.00 

150 

12 

7,750.00 

250 

14 

21,500.00 

150 

14 

8,275.00 

250 

16 

24,250.00 

244 


STEAM  POWER  PLANT  ENGINEERING 


TABLE  31. 

PROPORTIONS    OF    CHIMNEYS    FOR    FACTORY    STEAM    BOILERS,    COLLECTED 
FROM    PRACTICE.      (Button.) 


Height  of 

Internal  Dimensions. 

Ratio  of 

Thickness  of  Walls. 

above  the 

Top. 

Thickness 

Ground  in 

Feet. 

Size  of  Base  at  the 
Ground  Line. 

Size  of  Top. 

Internal 
Area. 

at  Base  in 
Inches  at 
Ground 

Thickness 
at  the  Top 
in  Inches. 

Line. 

40 

2'       6" 

1'     9*  sq. 

2.04 

18 

9 

60 

2'     11" 

2'     0"  sq. 

2.12 

18 

9 

70 

3'       4* 

2'     3"sq. 

2.13 

23 

9 

80 

3'       8* 

2'     6"sq. 

2.18 

28 

9 

90 

4'       0* 

2'     9'sq. 

2.27 

28 

9 

100 

4'       8* 

3'     O'diam. 

2.40 

28 

9 

110 

4'     10* 

3'     3*  diam. 

2.33 

28 

9 

120 

5'       Q" 

3'     6*  diam. 

2.40 

28 

9 

135 

6'      0" 

4'     0*  diam. 

2.30 

28 

9 

150 

4'       6" 

3'    0"  diam. 

2.25 

28 

14 

155 

6'      0* 

4'     6*  diam. 

1.78 

56 

14 

160 

9'      0* 

5'    O'sq. 

3.24 

36 

14 

170 

7'      6* 

5'    O'diam. 

2.25 

36 

14 

180 

6'      4" 

4'     6"  diam. 

2.00 

54 

14 

200 

5'      y 

3'     6*  diam. 

2.28 

36 

14 

225 

16'      0" 

6'     6"sq. 

4.00 

36 

14 

250 

19'       0* 

13'    O'diam. 

2.13 

40 

14 

300 

14'       0* 

9'    O'diam. 

2.42 

48 

14 

450 

21'       6* 

10'     2"  diam. 

4.35 

59 

14 

CHAPTER   VIII. 

MECHANICAL  DRAFT. 

147.  General.  —  The  intensity  of  natural  draft  in  a  chimney  depends 
mainly  upon  the  height  of  the  stack  and  the  temperature  of  the  chim- 
ney gases,  and  the  chimney  should  be  designed  to  meet  the  maximum 
requirements,  permitting  the  damper  to  be  partly  shut  at  times.    There 
is  usually  no  practicable  means  of  increasing  the  draft  after  the  maxi- 
mum has  been  reached.     Again,  chimney  draft  is  peculiarly  susceptible 
to  atmospheric  influence  and  may  be    seriously  impaired  by  adverse 
winds  and  air  currents.     Notwithstanding  these  apparent  limitations, 
by  far  the  greater  number  of  steam  power  plants  depend  upon  chim- 
neys for  draft,  and  for  obvious  reasons  as  will  be  discussed  later.     In 
many  cases  artificial  draft  has  a  great  advantage  and  under  certain 
conditions   is  indispensable;    it  is  very  flexible  and    readily  adjusted 
to  effect  various  rates  of  combustion,  irrespective  of  climatic  influ- 
ences, and  permits  any  degree  of  overload  without  undue  expenditure 
of  energy. 

Artificial  draft  may  be  broadly  classified  under  two  heads 

1.  The  vacuum  or  induced  draft  and 

2.  The  plenum  or  forced-draft  method. 

In  the  former  a  partial  vacuum  is  produced  above  the  fire  by  suitable 
apparatus,  and  the  effect  is  substantially  that  of  natural  draft. 

In  the  forced-draft  system  pressure  is  produced  in  the  ash  pit,  the 
air  being  forced  through  the  grate. 

In  both  systems  the  artificial  draft  is  usually  produced  by  either 

1.  Steam  jets  or 

2.  Centrifugal  fans  or  exhausters. 

148.  Steam  Jets.  —  Fig.  125  shows  an  application  of  a  ring  jet  to 
the  base  of  a  stack.     The  apparatus  is  very  simple,  inexpensive,  and 
easily  applied.     It  consists  essentially  of  a  ring  or  a  series  of  concen- 
tric rings  of  1-inch  or  IJ-inch  pipe,  perforated  on  the  upper  side  with 
T*fr  or  J-  inch  holes,  and  placed  in  the  base  of  the  stack,  so  that  the 
jets  are  discharged  upward,  thus  creating  a  draft  independent  of  the 
temperature  of  the  flue  gases.     The  steam  connection  to  the  jet  is 
generally  made  direct  to  the  boiler  and  not  to  the  steam  main,  though 
the  jet  is  often  produced  by  exhaust  steam. 

245 


246 


STEAM   POWER  PLANT  ENGINEERING 


Fig.  126  illustrates  a  Bloomsburg  jet,  which  involves  to  some  extent 
the  principle  of  the  ejector. 

The  increase  in  draft  produced  by  these  devices  as  ordinarily  installed 


FIG.  125.    Ring  Steam  Jet. 


* ~ 

FIG.  126.    Bloomsburg  Jet. 


is  not  great,  although  in  locomotive  practice  where  the  entire  exhaust 
is  discharged  up  the  stack  an  intense  draft  is  obtained. 

Fig.   127  shows  the  application  of  a  "  McClaves  argand   blower." 


FIG.   127.    McClaves  Argand  Blower. 

The  steam  is  discharged  below  the  grate  through  a  perforated  hollow 
ring,  as  indicated,  drawing  the  air  through  the  funnel  by  inspiration. 
This  creates  a  powerful  draft  by  forming  an  air  pressure  in  the  ash  pit, 


MECHANICAL  DRAFT 


247 


and  is  an  especially  useful  system  of  forcing  fires  for  boilers  which  need 
forcing  for  short  periods  only. 

Steam  jets  are  very  uneconomical,  since  a  large  amount  of  steam  is 
required  to  produce  good  results.  Table  32,  based  on  experiments  at 
the  New  York  Navy  Yard  to  determine  the  best  form  of  steam  jet 
for  producing  draft  in  launch  boilers,  shows  steam  consumptions  of 
8.3  to  21.2  per  cent  of  the  total  steam  made.  Table  33  gives  the  steam 
consumption  of  a  number  of  types  of  steam  jet  blowers. as  determined 
by  A.  J.  Whitham.  The  best  performance  is  4.6  per  cent  and  the 
poorest  11.1  per  cent  of  the  total  boiler  steam  generated.  Steam  jets 
below  the  grate  are  said  to  prevent  clinker  from  forming  where  fine 
anthracite  coals  are  used,  and  thus  to  assist  in  keeping  the  fire  free 
and  open. 

Steam  jets  arranged  above  the  grate  and  discharging  either  from  the 
side  walls,  front  wall,  or  bridge  wall,  oftentimes  assist  complete  com- 
bustion by  stirring  up  the  volatile  gases  and  air  and  insuring  a 
thorough  mixture,  thus  affording  one  of  the  simplest  and  frequently  a 
very  efficient  means  of  furthering  smokeless  combustion.  The  action 
is  of  course  purely  mechanical,  the  steam  in  itself  not  being  a  sup- 
porter of  combustion;  hence  if  the  air  supply  is  deficient  the  steam  jet 
is  of  no  avail  unless  arranged  to  carry  sufficient  air  along  with  the 
steam. 

Fig.  128  illustrates  such  an  application  to  a  hollow  bridge  wall. 
The  top  of  the  wall  is  fitted  with  a  small  cast-iron  column  M,  partially 


FIG.  128.    Application  of  Steam  Jets  to  Hollow  Bridge  Wall. 

imbedded  in  the  brickwork.  A  series  of  1-inch  holes  "  00,"  drilled 
near  the  top  of  the  casting,  furnish  exits  for  the  steam  and  air.  A 
steam  jet  in  one  end  of  the  column  induces  air  into  the  iron  chamber 
and  forces  it  across  the  fire  in  fine  streams.  Excessive  air  dilution  is 
avoided  by  partially  closing  the  ash-pit  doors  and  by  regulating  the 
intensity  of  the  jets.  An  installation  of  this  type  is  especially  effective 
in  connection  with  coal  having  a  tendency  to  fuse  and  seal  the  air  pas- 


248 


STEAM   POWER   PLANT  ENGINEERING 


sages  in  the  grate.  Two  Stirling  boilers  at  the  Armour  Institute  of 
Technology  equipped  with  this  device  gave  practically  smokeless 
combustion  at  all  normal  loads,  though  at  heavy  overloads  it  was 
sometimes  necessary  to  slightly  open  the  fire  doors.  Without  the  use 
of  the  jets  smoke  could  not  be  prevented  even  at  light  loads.  Analysis 
of  the  flue  gases  showed  but  a  slight  decrease  in  the  percentage  of  CO2. 

TABLE   32. 
RESULTS  OF  EXPERIMENTS  UPON  STEAM  JETS  AT  NEW  YORK  NAVY  YARD.* 


Index  of  Jet. 

Pounds  of  Water  Evaporated  per  Hour. 

A 

B 

C 

D 

E 

In  boiler  making  steam  
In  boiler  supplying  jets. 

463.8 
97.5 

21.2 

580.0 
120 

20.7 

361.25 
30 

8.3 

528.5 
63.2 

12.0 

545.00 
76.25 

19.0 

Per  cent  of  steam  used 
bv  iet..  . 

*  Annual  Report  of  the  Chief  of  the  Bureau  of  Steam  Engineering,  U.  S.  Navy,  1890. 


TABLE  33. 
CONSUMPTION    OF   STEAM    BLASTS    COMPARED,  f 


Per  Cent  of  Total 

Coal. 

Name  of  Blower. 

Per  Cent  of  Air 
Openings  in 
Gr&tc 

Pounds  of  Dry 
Coal  burned  per 
Hour  per  Square 

Steam  Generated 
in  the  Boilers 
that  is  required 

Foot  of  Grate. 

to  operate  the 

Steam  Blasts. 

Rice  

Young  

11 

25.8 

11.1 

Do  

.  .  .do  

11 

17.9 

7.0 

Do  

Wilkinson. 

7 

27.0 

10.8 

Buckwheat  

11 

27.3 

10.8 

Do  

do  

11 

16.7 

4.6 

Do  

....do  

26 

31.4 

8.9 

Do  

McClave  

11 

16.4 

6.7 

Do  

....do  

11 

26.1 

9.3 

Do  

Wilkinson  

7 

32.5 

7.8 

Do  

....do  

7 

45.4 

10.2 

t  Trans.  A.S.M.E..  Vol.  XVII.  — Whitham. 

149.  Parson  Smokeless  Furnace.  —  The  Parson  forced-draft  system 
for  smokeless  combustion,  applied  to  a  return  tubular  boiler  as  illus- 
trated in  Fig.  129,  comprises  a  specially  designed  grate  G,  depending 
upon  a  steam  jet  blower  A  for  draft.  Part  of  the  steam  is  admitted 


MECHANICAL  DRAFT 


249 


below  the  grate  and  part  over  the  fire  through  the  hollow  bridge  wall  H. 
The  supply  of  air  above  the  grate  is  regulated  by  means  of  damper  F. 
The  steam  to  blower  A  is  automatically  adjusted  by  regulator  N, 
which  is  actuated  by  the  steam  pressure.  The  steam  to  the  jet  is 
superheated  by  passing  the  supply  pipe  through  the  setting  as  indi- 


FIG.  129.    Parson's  Smokeless  Furnace: 

cated.     The  bridge  wall  H  is  provided  with  an  extension  platform  M 
for  holding  the  unburned  fuel  when  cleaning  the  fire. 

150.  Heinrich  Smokeless  Furnace.  —  Figs.  130  and  131  show  the  appli- 
cation of  the  Heinrich  system  of  forced  draft  to  a  return  tubular  boiler. 
Hot  air  is  taken  from  the  boiler  room  above  the  boilers  by  a  steam  jet 
blower  at  A  and  forced  into  the  superheating  chamber   below   the 
combustion  chamber.     From  this  chamber  part  of  the  air  is  drawn  by 
the  auxiliary  blowers  C  and  forced  through  tuyeres  above  the  grate, 
the  rest  passing  through  an  opening  beneath  the  bridge  wall  into  the 
ash  pit  and  up  through  the  bed  of  fuel.     Steam  for  the  blower  A  and 
the  auxiliaries  C  is  supplied  through  an  automatic  regulator  R,  which 
opens  when  the  steam  pressure  falls  below  the  required  value.     The 
manufacturers    (Heinrich  Manufacturing  Company,  Milwaukee,  Wis.) 
sell  this  apparatus  with  a  guarantee  of  15  per  cent  saving  in  fuel  over 
natural  draft,  common  grate  bars,  and  hand  firing. 

151.  Fan  Draft.  —  Fig.  132  shows  a  typical  installation  of  a  centrif- 
ugal fan  on  the  forced-draft  or  plenum  principle,  the  fan  creating  a 
pressure  in  the  ash  pit  and  forcing  air  through  the  fuel.     The  most 
approved  method  is  to  pass  the  air  through  the  bridge  wall,  thence 
toward  the  front  of  the  grate,  though  it  may  enter  through  an  under- 
ground duct  or  through  the  side  of  the  setting.     Forced  draft  is  usually 
adopted  in  old  plants  where  increased  demands  for  power  require  that 
the  boilers  be  forced  far  above  their  rating  to  save  the  heavy  expense 


250 


STEAM   POWER   PLANT   ENGINEERING 


of  new  boilers,  or  in  plants  burning  refuse,  anthracite  culm  or  screen- 
ings, which  require  an  intense  draft  for  efficient  combustion.  Forced 
draft  is  also  well  adapted  for  underfeed  stokers  of  the  retort  type, 
hollow  blast  grates,  and  the  closed  fire  hole  system.  The  air  supply 
may  be  taken  from  an  air  chamber  built  around  the  breeching,  thereby 
supplying  the  heated  air  to  the  fan  and  effecting  a  lower  temperature 
in  the  breeching  and  a  higher  temperature  in  the  furnace.  The 
objection  is  sometimes  raised  against  forced  draft  that  the  gases  tend 


FIG.  130.   Heinrich  Smokeless  Furnace  (Sectional  Elevation). 


FIG.  131.  Heinrich  Smokeless  Furnace  (Sectional  Plan). 

to  pass  outward  through  the  fire  door  when  the  fire  is  cleaned  or  re- 
plenished, since  the  pressure  in  the  furnace  is  greater  than  atmospheric. 
This  objection  may  usually  be  overcome  by  suitable  dampers  in  the 
blast  pipe  which  are  closed  on  opening  the  fire  doors.  With  a  boiler 
plant  of  1000  horse  power  or  more  the  cost  of  a  forced-draft  fan, 
engine,  and  stack  will  approximate  20  to  30  per  cent  of  the  outlay  of 
an  equivalent  brick  chimney.  The  power  consumption  will  depend 


MECHANICAL  DRAFT 


251 


upon  the  character  and  efficiency  of  the  motor  or  engine  and  will  range 
from  1  to  5  per  cent  of  the  total  capacity. 

Induced  draft  as  illustrated  in  Fig.  133  is  perhaps  the  most  com- 
mon substitute  for  natural  draft  and  is  extensively  used  in  street  rail- 
way and  lighting  plants  which  have  high  peak  loads,  being  ordinarily 
installed  in  connection  with  fuel  economizers.  The  suction  side  of  the 
fan  is  connected  with  the  uptake  or  breeching  of  the  boiler  or  bat- 
teries of  boilers  and  the  products  of  combustion  usually  exhausted 
through  a  stub  stack.  The  illustration  shows  a  typical  installation  in 
which  two  fans  of  the  duplex  type  are  placed  above  the  boiler  setting. 
The  fan  ducts  are  generally  designed  with  a  by-pass  direct  to  the  stack 
to  be  used  in  case  of  accident  or  when  mechanical  draft  is  not  required. 


FIG.  132.   Typical  Forced-Draft  System. 


Since  the  fan  handles  hot  gases  it  must,  under  the  ordinary  con- 
ditions of  practice,  have  a  capacity  approximately  double  that  of  a 
forced-draft  fan  delivering  cold  air,  but  the  gases  being  of  lower 
density  the  power  required  per  cubic  foot  moved  is  less. 

With  forced  draft  about  300  cubic  feet  of  air  are  required  per  pound 
of  coal;  with  induced  draft  the  fan  must  handle  twice  this  volume  if 
the  gases  are  exhausted  at  500  degrees  F.  or  450  cubic  feet  if  exhausted 
at  300  degrees  F.,  a  temperature  to  be  expected  in  connection  with 
economizers. 

The  advantages  of  induced  draft  over  forced  draft  are  very  pro- 
nounced. The  pressure  in  the  furnace  is  less  than  atmospheric,  there- 
fore it  is  not  necessary  to  shut  off  the  draft  in  cleaning  fires  or  ash  pit, 
and  the  fire  burns  more  evenly  over  the  entire  grate  area  and  requires 
less  attention  than  with  forced  draft.  An  induced-draft  plant  costs 
considerably  more  than  forced  draft  on  account  of  the  larger  fan 


252 


STEAM  POWER  PLANT  ENGINEERING 


required,  but  the  operating  expenses  are  but  little  greater.  With  a 
boiler  plant  of  1000  horse  power  or  more  the  cost  of  a  single  induced- 
draft  fan,  engine,  stack,  etc.,  will  approximate  40  to  50  per  cent  of  the 
outlay  required  for  a  brick  chimney  of  equivalent  capacity,  and  the 
double-fan  outfit  will  approximate  50  to  60  per  cent.  The  double-fan 


FIG.  133.   Typical  Induced-Draft  System. 

system  is  particularly  adapted  to  plants  which  operate  continuously 
and  where  even  a  temporary  break-down  is  a  serious  inconvenience. 

Advantages  of  Mechanical  Draft:  Am.  Elecn.,  June,  1898,  p.  244,  Feb.,  1902, 
p.  63;  Eng.  Rev.,  Sept.,  1901,  p.  4;  Eng.  Mag.,  April,  1901,  p.  81,  March,  1900, 
p.  931;  Elec.  Rev.,  Lond.,  Feb.  3,  1899,  p.  186;  Cassier's  Mag.,  Nov.,  1898,  p.  48, 
Jan.,  1905,  p.  252,  March,  1906;  St.  Ry.  Rev.,  July  15,  1901,  p.  415;  Engr.  U.S., 
July  16,  1906,  p.  475;  Elec.  Eng.,  Aug.  11,  1905,  p.  193. 

Application  of  Mechanical  Draft  to  Stationary  Boilers  :  Power,  Dec.,  1900,  p.  30; 
West.  Elecn.,  Feb.  16,  1901,  p.  118;  Jour.  West.  Soc.  Engrs.,  March  19, 1902,  p.  271 ; 
St.  Ry.  Rev.,  July,  1899,  p.  463;  Cassier's  Mag.,  Nov.,  1898,  p.  48;  Elec.  Rev.,  July 
27,  1898,  p.  52;  Engr.  U.S.,  Jan.  1,  1907. 

152.  Performance  of  Fans.  —  The  first  satisfactory  theory  of  centrif- 
ugal fans  was  promulgated  by  Daniel  Murgue  in  1872.  He  proved 


MECHANICAL  DRAFT 


253 


that  theoretically  the  maximum  pressure  created  by  a  perfect  fan  is 
equivalent  to  twice  the  head  which  would  produce  a  velocity  equal  to 
that  of  the  periphery.  Thus 

H  =^~,  (60) 

in  which 

H  =  maximum  difference  in  pressure  in  feet  of  air, 
u  =  peripheral  velocity  in  feet  per  second,  and 
g  =  acceleration  of  gravity  32.2. 

A  and  5,  Fig.  134,  represent  Pitot  tubes  inserted  in  the  discharge 
pipe  of  a  centrifugal  blower,  A  being  bent  to  face  the  current,  while  B 
is  at  right  angles  to  it.  A  receives  the  full  impact  of  the  stream,  and 


(A) 


n 

J 


(B) 


ORIFICE     CLOSCD 


FIG.  134. 

the  manometer  indicates  the  total  pressure,  static  and  velocity,  while 
B  registers  the  static  pressure  only.  With  the  discharge  orifice  closed, 
as  in  Fig.  134,  the  velocity  becomes  zero,  and  the  water  depression  in 
both  manometers  will  be  the  same,  due  to  the  static  pressure,  which, 
according  to  Murgue's  theory,  will  be  a  maximum  and,  ignoring  fric- 

u2 
tion  or  eddy  currents  = 

9 

Example  :  Determine  the  maximum  pressure,  in  inches  of  water, 
which  a  perfect  fan  would  exert  with  discharge  orifice  closed;  diameter 
of  fan  6  feet;  r.p.m.  318. 

The  peripheral  velocity  is 

u  =  2  nrn  =  6.28  X  3  X  318  =  6000  feet  per  minute. 

=  100  feet  per  second. 

Substituting  in  Murgue's  formula, 

u  =  100  and  g  =  32.2, 

H  =     =  31° feetj 


254 


STEAM   POWER  PLANT  ENGINEERING 


i.e.,  the  pressure  created  by  the  fan  would  be  equivalent  to  the  weight 
of  a  column  of  air  310  feet  high,  or,  assuming  an  air  temperature  of 
75  degrees  F.,  an  equivalent  head  in  inches  of  water  of 
310X0.074495 


144  X  0.0361 


=  4.45  inches. 


(0.074495  =  density  of  air  at  75  degrees  F.  and  0.0361  =  pressure  pro- 
duced by  one  inch  of  water  in  pounds  per  square  inch.) 

If  the  discharge  orifice  be  opened  to  its  maximum  (Fig.  135)  the  static 
pressure   indicated   by  manometer  B  becomes  zero,  since  there  is  no 


(A) 


(B) 


ORIFICE    WIDE  OPEN 


FIG.  135. 

resistance  due   to   the   air  flow,  while   the  water  in  A  stands   at   a 
height  H  the  exact  equivalent  of  the  velocity  head  in  accordance  with 

the  hydraulic  formula,  

v  =  V2gH, 

in  which  v  is  the  velocity  of  the  air  in  feet  per  second. 

If  the  orifice  be  partially  closed,  say  50  per  cent,  as  in  Fig.  136,  B  indi- 
cates the  static  pressure,  while  A  gives  the  dynamic  or  total  pressure 
due  to  both  velocity  and  resistance.  The  difference  between  A  and  B 
is  therefore  the  pressure  due  to  velocity  alone.  By  connecting  the 
two  manometers  as  indicated  in  Fig.  136  C  the  velocity  pressure  is 
given  directly. 

PRESSURE.  —  According  to  Murgue's  theory  the  maximum  pressure 
which  may  be  developed  by  a  blower  or  exhauster  varies  with  the  square 
of  the  speed  and  may  be  expressed 

Cdu* 
p-—, 

in  which 

p  =  pressure,  pounds  per  square  foot. 
d  =  density  of  the  air,  pounds  per  cubic  foot. 
u  =  peripheral  velocity,  feet  per  second. 
C  =  a  coefficient  obtained  by  experiment. 


MECHANICAL    DRAFT 


255 


Tables  34  and  35  give  the  relationship  between  pressure  and  speed 
for  various  sizes  of  forced  and  induced-draft  fans. 

Fig.  139  shows  the  relationship  between  pressure  and  speed  in  a 
45-inch  Buffalo  blower  as  tested  at  the  Armour  Institute  of  Technology. 

VELOCITY  OF  DISCHARGE.  —  The  maximum  velocity  of  the  air  leav- 
ing the  tips  of  the  blades  varies  directly  as  the  peripheral  speed, 


V  =  Ku, 


(63) 


in  which 


V  =  velocity  of  the  air  discharged,  feet  per  second. 
K  =  a  coefficient  obtained  by  experiment. 
u  =  peripheral  velocity,  feet  per  second. 


H 

s^ 

r 
S 

1 

T~ 
h 
_1 

«=i^ 

- 

7 

^ 

T 
H-h 

_L 

A  ~~       *"    r*> 

(A) 

I 

(B) 

r> 

(C) 

/ 

ORIFICE    PARTLY    CLOSED 


FIG.  136. 


For  practical  purposes  the  velocity  of  discharge  with  outlet  wide  open 
may  be  assumed  to  be  that  of  the  periphery. 

CAPACITY.  —  The  relationship  between  capacity  and  speed,  capacity 
and  discharge  opening  for  a  45-inch  pressure  blower  is  given  in  Figs.  139 
and  140. 

As  will  be  noted,  the  capacity  varies  almost  directly  with  the  speed  of 
the  wheel  and  the  area  of  discharge  as  expressed  by  the  equation 


Q  =  BnADN, 


(63a) 


in  which 


Q  =  cubic  feet  discharge  per  minute. 

B  =  coefficient  determined  from  experiment. 

A  =  area  discharge  opening,  square  feet. 

D  =  diameter  of  the  wheel. 

N  =  r.p.m.  of  the  wheel. 


256  STEAM   POWER  PLANT  ENGINEERING 

POWER.  —  The  power  required  to  drive  a  fan  is  proportional  to  the 
cube  of  the  speed, 

Horse  Power  =  XAN3,  (64) 

in  which 

X  =  a  coefficient  determined  by  experiment. 
A  =  area  discharge  outlet,  square  feet. 
N  =  r.p.m. 

The  marked  increase  in  power  required  for  even  a  moderate  increase 
in  speed  should  be  borne  in  mind  in  selecting  a  fan.  It  is  as  a  rule 
more  economical  to  err  in  selecting  too  large  a  fan  than  one  which  must 
be  forced  above  its  rated  capacity. 

In  practice  the  size  of  fan  is  proportioned  upon  experience  rather 
than  theory,  the  usual  procedure  necessitating  the  use  of  curves  based 
upon  the  performance  of  fans  of  the  type  under  consideration. 

The  curves  in  Fig.  137  were  computed  by  Mr.  F.  R.  Still  of  the 
American  Blower  Company,  and  give  the  performance  of  steel-plate 
fans  as  manufactured  by  this  company.  These  curves  apply  to  this 
type  and  make  of  fan  only,  though  the  difference  is  not  very  great 
for  any  type  of  centrifugal  fan.  The  "  ratio  of  opening  "  refers  to  the 
actual  percentage  of  opening  compared  with  the  total  discharge.  The 
"  ratio  of  effect  "  is  the  relative  effect  produced  by  restricting  the 
discharge.  The  abbreviations  are  as  follows  : 

D.P.  =  dynamic  or  total  pressure. 
P.V.P.  =  pressure  created  by  a  column  of  air  moving  at  the 

same  velocity  as  the  periphery. 
S.P.  =  static  pressure. 
V.P.  =  velocity  pressure  =  D.P  -  S.P. 

Suppose  a  fan  with  an  unrestricted  inlet  and  outlet  delivers  25,000 
cubic  feet  of  air  per  minute  against  a  head  (D.P.)  of  0.33  inch  with  a 
peripheral  velocity  requiring  6.16  horse  power.  It  appears  from  the 
curves  that  if  the  discharge  outlet  is  restricted  to  50  per  cent  of  the 
full  area,  only  12,500  cubic  feet  will  be  delivered;  the  pressure  will  be 
increased  to  1.03  inches,  and  the  power  required  drops  to  4.84  horse 
power.  If  the  outlet  be  still  further  reduced  to  20  per  cent  of  the  full 
opening  the  capacity  will  drop  to  5000  cubic  feet,  the  pressure  will 
increase  to  1.15  inches,  and  the  power  will  be  decreased  to  3.45  horse 
power.  With  a  discharge  area  of  60  per  cent,  the  mechanical  efficiency 
is  a  maximum,  and  equal  to  about  43  per  cent.  With  orifice  closed 
the  horse  power  required  to  drive  the  fan  is  about  37  per  cent  of  that 
required  when  discharging  the  maximum  volume  of  air. 


MECHANICAL  DRAFT 


257 


0         10        20        30        40        50        60        70        80        90       100      110       130      130      140 
Ratio  of  Effect  Per  Cent 

FIG.  137.    Performance  of  Steel  Plate  Fans. 


-,» 


i 

- 


\ 


45  In 

Buffalo  Blower 
Speed  Constant 

1500  R.P.M. 
Discharge  Area  Variable 


.1  0.2  03  0.4 

Area.of  Discharge  Opening,  Square  Feet 

FIG.  138. 


258 


STEAM   POWER  PLANT  ENGINEERING 


Curve  "  K  "  in  Fig.  137  was  determined  from  the  empirical  formula 
(based  upon  Murgue's  theorem) 

(65) 


in  which 

A  =  area  of  the  inlet  orifice,  square  feet. 

Q  =  volume  of  gas,  thousands  of  cubic  feet  per  minute. 

P  =  draft  at  the  inlet  in  inches  of  water. 

K  =  constant  determined  by  experiment. 


45  In. 

Buffalo  Blower 

Discharge  Area  Constant 

Speed  Variable 


700 


800         900         1000        1100        1200 
Revolutions  per  Minute 

FIG.  139. 


1300       1400       1500 


The  curves  in  Figs.  138  and  139  are  plotted  from  tests  made  at  the 
Armour  Institute  of  Technology  on  a  45-inch  Buffalo  pressure  blower, 
and  are  characteristic  of  this  type  of  fan. 

Theory  of  Fans  :  Power,  May,  1907,  p.  287;  Engr.,  Oct.  9,  1903,  p.  512;  Mach., 
Aug.,  1898;  Sib.  Jour,  of  Eng.,  Nov.,  1902 ;  Heat  and  Vent.,  Jan.  15, 1897,  July,  1899; 
Prac.  Engr.,  Jan.  16,  1903;  Mech.  Engr.,  April  18,  1903;  Eng.  Rec.,  Oct.  11,  1902. 

Pressure  Fans  vs.  Exhaust  Fans:  Bulletin  Am.  Inst.  Min.  Engrs.,  Feb.,  1909. 

153.  Determination  of  Size  of  Fan. —  The  following  analysis,  based 
upon  a  paper  on  Mechanical  Draft  by  F.  R.  Still  of  the  American 


MECHANICAL  DRAFT  259 

Blower  Company,  gives  a  good  idea  of  the  usual  procedure  in  deter- 
mining the  size  of  fan  for  an  induced  draft  installation.  (Jour.  West. 
Soc.  Engr.,  May,  1902.) 

Example  :  Determine  the  size  of  induced  fan  and  the  approximate 
power  required  to  drive  it,  for  a  boiler  plant  rated  at  1000  horse 
power;  temperature  of  flue  gases  500  degrees  F.;  heat  value  of  coal 
14,000  B.T.U.  per  pound;  ash  5  per  cent;  draft  required,  1  inch  of 
water  pressure. 

Assuming  a  boiler  efficiency  of  70  per  cent,  the  evaporation  will  be 
14,00px  a70  =  1Q  15  pounds  Of  water  from  and  at  212  degrees  F. 


per  pound  of  coal. 

Since  one  boiler  horse  power  is  equivalent  to  the  evaporation  of 
34.5  pounds  of  water  per  hour  from  and  at  212  degrees  F.,  the  evapora- 
tion per  hour  will  be  34.5  X  1000  =  34,500  pounds,  and  the  coal 
burned  per  hour, 

34>5QO  =  3400  pounds. 

Allowing  18  pounds  of  flue  gas  per  pound  of  combustible,  5  per  cent 
for  ash  and  5  per  cent  for  leaks,  the  fan  will  have  to  handle,  at  500 
degrees  F.,  approximately  20  X  3400  =  68,000  pounds  of  gas  per 
hour,  or  26,000  cubic  feet  per  minute.  It  is  customary,  when  little  is 
known  about  a  plant  in  which  a  fan  is  to  be  installed,  to  assume  that 
the  resistance  is  equivalent  to  restricting  the  discharge  outlet  25  per 
cent.  Hence  in  this  problem  the  various  factors  are  referred  to  a 
"  ratio  of  opening  "  of  75  per  cent  (see  Fig.  137). 

From  formula  65,  the  area  of  the  inlet  should  be 

KQ      0.485  X  26  - 

A  =  —  —  =  -  —  -  =  12.6  square  feet, 

which  corresponds  to  a  diameter  of  48  inches.     (K  =  0.485  is  taken 
from  the  curves  in  Fig.  127.) 

The  area  of  the  inlet  should  not  exceed  40  per  cent  of  the  area  of  the 
side  of  the  wheel;  the  latter,  then,  will  be 

—  -1-  =31.5  square  feet, 
0.4 

which  corresponds  to  a  diameter  of  76  inches  (6.3  feet). 

Referring  to  Fig.  137,  the  ratio  of  dynamic  pressure  to  peripheral 
velocity  pressure  (D.P.  to  P.V.P.  at  75  per  cent  opening)  is  0.73.  The 

peripheral  velocity  pressure  will  be  —  —  =  1.37  inches  of  water. 

0.73 


260  STEAM  POWER  PLANT  ENGINEERING 

The  peripheral  velocity  is 

U  =  V2gH'  =  8.03 VF7,    where     R'    is    the    peripheral    velocity 
pressure  expressed  in  feet  of  gas,  or, 

Since  R9  =     ^  * — _  ,  where  p  =  inches  water, 

0.0478  X  12 

17=    87.5  VT37. 

=  102.5  feet  per  second. 
=  6150  feet  per  minute. 

The  maximum  effective  discharge  area  which  an  inclosed  fan  of  this 
type  may  have,  and  still  maintain  the  pressure  equivalent  of  the  per- 
ipheral velocity,  is  usually  called  the  "  blast  area."  With  a  larger  area 
the  pressure  will  be  reduced,  but  with  a  smaller  area  will  remain  sub- 
stantially constant.  The  velocity  of  the  discharge  is  practically  that 
of  the  tips  of  the  blades,  whence  the  blast  area  is  equal  to 

26  000 

'        =  4.23  square  feet,  which,  with  this  type  of  fan  is  found  to  be 
oloO 

about  J  the  projected  rectangle  of  the  wheel,  therefore, 
The  projected  rectangle  =  4.23  X  3  =  12.7  square  feet. 
The  proper  width  of  periphery  is  found  by  dividing  this  area  by  the 

wheel  diameter,  thus, 

12  7 

width  of  blades  =  ==i  =  2.02  feet  =  24.2  inches, 
6.3 

and  speed  of  fan  = =  311  r.p.m. 

3.14  X  6.3 

PP  __  Volume  of  gas  (cu.  ft.  per  min. )  X  Pressure  (Ib.  per  sq.  ft. ) 
33,000  X  efficiency  of  fan 

W  =  26'000  X  5-2  =  10.2  brake  horse  power. 
33,000  X  0.4 

(5.2  =  pressure  in  pounds  per  square  foot  equivalent  to  one  inch  of 
water,  and  0.4  is  the  mechanical  efficiency  for  75  per  cent  opening  as 
taken  from  curve  in  Fig.  137.) 

Assuming  a  steam  consumption  of  70  pounds  per  brake  horse  power 
for  a  small,  simple  non-condensing  high-speed  engine,  the  steam  con- 
sumed per  hour  will  be 

10.2  X  70  =  714  pounds  per  hour,  or  2.3  per  cent  of  the  total  steam 
capacity  of  the  boilers. 

Table  34  gives  the  capacity  and  horse  power  required  for  various 
sizes  of  forced-draft  fans,  and  Table  35  gives  similar  data  for  induced- 
draft  fans. 


MECHANICAL  DRAFT 


261 


154.  Chimney  vs.  Mechanical  Draft.  —  The  choice  of  chimney  or 
mechanical  draft  depends  largely  upon  local  conditions.  Many  power 
plants  with  tall  stacks  are  provided  with  forced-draft  apparatus  to  be 
used  in  emergencies,  but  as  a  general  rule  where  ordinances  require 
high  chimneys  mechanical  draft  is  not  considered.  In  a  few  isolated 
cases  stokers  of  the  forced-draft  type  are  used  in  connection  with 
chimneys  as  high  as  250  feet,  but  such  installations  are  rare,  and  not 
to  be  recommended. 

Where  there  are  no  limitations  to  the  height  of  stack,  mechanical 
draft  offers  many  advantages  over  chimney  draft,  especially  for  rail- 
road work  and  large  lighting  plants.  With  certain  types  of  grates  and 
for  low-grade  fuels  and  anthracite  culm  or  dust,  it  is  indispensable. 
Again,  where  a  fair  quality  of  fuel  is  obtainable  the  size  of  plant  may 
determine  the  choice. 

First  Cost:  In  small  plants  of  say  100  to  150  horse  power  the  cost 
of  a  guyed  steel  chimney,  75  feet  in  height  or  less,  would  be  consider- 
ably less  than  that  of  a  mechanical-draft  system,  and  once  erected  cost 
practically  nothing  for  operation,  while  the  power  required  to  operate 
a  fan  in  so  small  a  plant  would  amount  to  5  per  cent  or  more  of  the 
total  steaming  capacity. 

TABLE  34. 

CAPACITIES    OF    FORCED-DRAFT    FANS. 
(Power.) 

For  Forced  Draft,  Temperature  of  Air  60°. 


Diam- 
eter of 
Fan. 

Cubic  Feet 
of  Air  De- 
livered to 
Furnace 

Pressure  in  Inches  of  Water. 

0.5 

0.75 

1.00 

1.25 

1.50 

2.00 

2.50 

per 

3j 

PH 

S 

PH 

Sti 

PH 

§ 

PH 

S 

PH 

9 

PH 

§ 

PH 

Minute. 

PH 

a 

PH 

a 

PH 

a 

PH 

a 

PH 

a 

PH 

a 

PH 

a 

PH 

(4 

PH 

pi 

PH 

pi 

p4 

2'    6* 

4,200 

510 

1.6 

560 

1.8 

600 

1.9 

640 

2.1 

710 

2.3 

780 

2.5 

850 

2.7 

3' 

5,800 

430 

2.2 

460 

2.4 

490 

2.6 

530 

2.8 

590 

3.1 

640 

3.4 

710 

3.8 

3'    6* 

7,800 

360 

3.0 

400 

3.3 

420 

3.5 

450 

3.8 

500 

4.2 

550 

4.6 

610 

5.1 

4' 

10,000 

320 

3.9 

350 

4.2 

370 

4.4 

400 

4.9 

440 

5.4 

480 

5.9 

530 

6.5 

4'    6* 

12,400 

290 

4.8 

310 

5.2 

330 

5.6 

360 

6.0 

400 

6.7 

430 

7.3 

470 

8.0 

5' 

15,200 

250 

5.9 

270 

6.4 

290 

6.8 

310 

7.4 

350 

8.2 

380 

8.9 

420 

9.8 

5'    6* 

18,200 

230 

7.0 

250 

7.7 

270 

8.2 

300 

8.8 

330 

9.8 

360 

10.6 

390 

11.8 

6' 

21,400 

210 

8.3 

230 

9.1 

250 

9.6 

260 

10.4 

290 

11.5 

320 

12.5 

350 

13.9 

V 

28,800 

180 

11.2 

200 

12.2 

210 

13.0 

230 

14.0 

250 

15.5 

280 

16.8 

300 

18.7 

8' 

37,200 

160 

14.4 

170 

15.7 

190 

16.7 

200 

18.1 

220 

20.1 

240 

21.8 

270 

22.5 

9' 

46,800 

140 

18.1 

160 

19.8 

170 

21.1 

180 

22.7 

200 

25.3 

220 

27.4 

240 

30.3 

10' 

57,400 

130 

22.2 

140 

24.3 

150 

25.8 

160 

27.9 

180 

3.1 

200 

33.6 

210 

37.2 

Discharge  velocity  2000  feet  per  minute. 


262 


STEAM  POWER  PLANT  ENGINEERING 


TABLE   35. 
CAPACITIES    OF    INDUCEI>-DRAFT    FANS. 

(Power.) 


For  Induced  Draft,  Temp,  of  Flue  Gases  500°. 


Cubic  Feet 
of  Air  at 

Pressure  in  Inches  of  Water. 

Diam- 

60°Temp. 

0.5 

0.75 

1.00 

1.25 

1.50 

2.00 

2.50 

eter  of 

Drawn 

Fan. 

into    Fur- 

• 

. 

. 

. 

. 

. 

nace  per 

^ 

PH' 

S 

PH' 

SB 

PH* 

W 

PH' 

S 

PH* 

H 

PH' 

B 

PH* 

Minute. 

pi 

M 

PH 
Pi 

PH' 

pi 

W 

PH 

pi 

W 

PH 
PH 

H 

PH' 

pi 

W 

PH 

pi 

a 

2'    6" 

3,000 

688 

2.2 

756 

2.4 

810 

2.6 

864 

2.8 

958 

3.1 

1053 

3.4 

1147 

3.6 

3' 

4,200 

580 

3.0 

621 

3.2 

661 

3.5 

715 

3.8 

796 

4.2 

864 

4.6 

958 

5.1 

3'    6" 

5,700 

486 

4.0 

540 

4.5 

567 

4.7 

607 

5.1 

675 

5.7 

742 

6.2 

823 

6.9 

4' 

7,300 

432 

5.3 

472 

5.7 

500 

6.1 

540 

6.6 

594 

7.3 

648 

8.0 

715 

8.8 

4'    6" 

9,300 

390 

6.5 

418 

7.0 

445 

7.5 

486 

8.1 

540 

9.0 

580 

9.8 

634 

10.8 

5' 

11,100 

337 

8.0 

364 

8.6 

391 

9.2 

418 

10.0 

472 

11.1 

513 

12.0 

567 

13.2 

5'    6* 

13,300 

310 

9.5 

337 

10.4 

364 

11.1 

405 

11.9 

445 

13.2 

486 

14.3 

526 

15.9 

6' 

15,600 

283 

11.2 

310 

12.3 

337 

13.0 

351 

14.0 

391 

15.5 

432 

16.9 

472 

18.7 

7' 

21,000 

243 

15.1 

270 

16.5 

283 

17.5 

310 

18.9 

337 

20.9 

378 

22.6 

405 

25.2 

8' 

27,100 

216 

19.4 

230 

21.2 

256 

22.5 

270 

24.4 

297 

27.1 

324 

29.4 

364 

30.4 

9' 

34,200 

189 

24.4 

216 

26.7 

230 

28.5 

243 

30.6 

270 

34.1 

297 

37.0 

324 

40.9 

10' 

41,900 

175 

30.0 

190 

32.8 

202 

34.8 

216 

37.6 

243 

41.8 

270 

45.3 

283 

50.2 

A  tall,  self-supporting  chimney  for  larger  plants,  however,  is  very 
costly  as  compared  with  a  fan  system  of  equal  capacity.  For  example, 
a  brick  chimney  175  feet  high  and  10  feet  in  diameter,  foundation 
and  all,  capable  of  furnishing  the  necessary  draft  for  a  3000-horse-power 
plant,  will  cost  about  $10,000.  A  two-fan  induced  system  of  equiv- 
alent capacity  will  cost  in  the  neighborhood  of  $5000,  a  one-fan 
system  $3500,  and  a  forced-draft  system  $2500.  See  Fig.  140.  With 
interest  at  5  per  cent,  depreciation  5  per  cent,  taxes  1  per  cent,  and 
insurance  one-half  per  cent,  the  annual  fixed  charges  will  be  $575, 
$402.50,  $287.50,  respectively,  for  the  fan  equipment. 

Depreciation  and  Maintenance  :  The  depreciation  of  a  well-designed 
masonry  or  concrete  stack  is  very  low,  and  2  per  cent  is  a  liberal  factor. 
Maintenance  is  practically  negligible,  as  it  requires  no  attention  what- 
ever for  years.  A  steel  stack,  however,  must  be  kept  well  painted  or 
corrosion  will  take  place  rapidly.  The  depreciation  and  maintenance 
charges  on  a  mechanical-draft  system  will  range  from  4  per  cent  to  10 
per  cent  of  the  original  outlay. 

Cost  of  Operation:  Once  erected,  the  comparative  cost  of  operating 
a  chimney  is  practically  nothing;  that  is,  of  course,  on  the  assumption 
that  the  chimney  and  fan  exhaust  equal  volumes  of  gas  per  pound  of 


MECHANICAL  DRAFT 

fuel  and  at  the  same  temperature.  A  fan  system  requires  for  its  opera- 
tion from  one  and  one-half  per  cent  to  five  per  cent  of  the  total  steaming 
capacity  of  the  plant,  depending  upon  the  type  and  character  of  the 
fan  engine  or  motor,  and  the  conditions  of  operation. 

Efficiency:  With  fan  draft  a  very  thick  fire  can  be  maintained  on 
the  grate,  thus  permitting  a  high  rate  of  combustion,  and  minimum 
air  per  pound  of  fuel,  both  of  which  result  in  increased  boiler  efficiency. 


1000 


2000 


Horse  Power 
FIG.  140.   Comparative  Costs  of  Chimneys  and  Mechanical  Draft.      (W.  B.  Snow.) 

The  influence  of  the  rate  of  combustion  on  air  supply  is  illustrated  in 
Fig.  141.  For  the  same  temperature  of  discharge  each  pound  of  air 
in  excess  of  theoretical  requirements  results  in  a  loss  of  about  one  per 
cent  of  the  total  heat  in  the  fuel.  (See  Table  3.)  With  fan  draft  an 
average  figure  is  18  pounds  of  air  per  pound  of  bituminous  coal  against 
24  pounds  for  the  chimney,  a  saving  of  5  per  cent  in  favor  of  the  fan. 
Again,  a  fan  permits  of  a  low  temperature  of  the  flue  gases  without 


264 


STEAM  POWER   PLANT  ENGINEERING 


affecting  the  draft,  while  lowering  the  temperature  in  the  chimney 
reduces  the  draft  as  shown  in  Table  24.  From  Table  4  we  see  that 
a  reduction  in  flue  gas  temperature  of  25  degrees  F.  will  increase  the 
boiler  efficiency  about  one  per  cent.  With  an  economizer  the  flue 
gases  may  be  reduced  to  350  degrees  F.;  with  a  net  saving  of  about 
500  —  350  =  150,  or  6  per  cent  of  the  total  fuel.  It  is  in  this  connection 
that  the  fan  draft  is  peculiarly  suitable.  Of  course,  the  chimney  may 
be  provided  with  an  economizer,  effecting  the  same  reduction  in  tem- 


10 


20  30 

ib.Coal  Burned  Per  Sq.Ft.Grate  Per  Hx. 


40 


FIG.  141 .    Influence  of  Rate  of  Combustion  on  Air  Supply.  —  Forced  Draft. 

perature,  but  its  height  must  be  made  sufficiently  great  to  overcome 
the  additional  resistance  of  the  economizer  and  the  reduction  in  tem- 
perature of  the  chimney  gases. 

Flexibility :  With  a  fan  the  draft  may  be  readily  regulated  for 
sudden  increased  or  decreased  requirements,  independent  of  the  boiler 
performance.  Damp  and  muggy  days  appreciably  affect  the  draft  of 
a  chimney,  as  do  adverse  air  currents  and  high  winds. 

Smoke:  Smokeless  combustion  is  more  readily  effected  with  arti- 
ficial draft  than  with  natural  draft,  as  a  thicker  fire  can  be  carried,  and 
the  correct  proportion  of  air  can  be  more  readily  adjusted. 

Comparative  Tests  of  Chimney  and  Mechanical  Draft:  Power,  July,  1901,  p.  22; 
Eng.  Rec.,  July  25,  1903,  p.  102;  Eng.,  U.  S.,  May  1,  1899,  p.  105;  Engr.,  U.  S., 
April  15,  1907. 

155.  Balanced  Draft.  —  Fig.  142  illustrates  an  application  of  the 
McLean  "  Balanced  Draft "  system  to  a  water-tube  boiler.  The 
equipment  consists  of  a  blower,  the  speed  of  which  is  regulated 


MECHANICAL   DRAFT 


265 


i        do 


266  STEAM  POWER  PLANT  ENGINEERING 

by  the  steam  pressure,  so  that  the  draft  in  the  fire  box  is  main- 
tained at  approximately  atmospheric  pressure.  The  chief  claims  for 
this  system  are  (1)  the  velocity  of  the  gases  over  the  tubes  is  reduced, 
and  short  circuiting  is  prevented;  (2)  the  correct  proportion  of  air  to 
fuel  is  readily  maintained;  (3)  infiltration  of  air  through  the  setting 
is  impossible,  as  the  pressures  are  "balanced";  (4)  sudden  changes 
in  load  are  correctly  taken  care  of.  Tests  of  the  apparatus  at  the 
Fuller  Building,  New  York,  gave  excellent  results  (Trans.  A.S.M.E., 
26-641). 


CHAPTER  IX. 

STEAM  ENGINES. 

156.  Introductory.  —  The  reciprocating  steam  engine  is    the  most 
widely  distributed  and  generally  adopted  prime  mover   in  the  power 
world  although  its  field  of  usefulness  has  been  greatly  encroached  upon 
in  recent  years  by  the  steam  turbine  and  the  gas  engine.     The  steam 
turbine  has  practically  superseded  the  piston  engine  for  large  steam 
electric  plants,  while  in  other  fields  the  gas  engine  offers  many  advan- 
tages, but  the  reciprocating  steam  engine  is  still  an  important  heat 
engine  and  will  probably  continue  to  be  a  factor  in  the  power  world 
for  years  to  come. 

The  type  of  engine  best  suited  for  a  given  installation  is  the  one 
which  delivers  the  required  power  at  the  lowest  cost,  measured  in 
dollars  and  cents,  taking  into  consideration  interest  on  the  investment, 
operating  expenses,  maintenance  and  depreciation. 

157.  Ideal  Engine.  —  The   thermal   efficiency   of  the  steam   engine 
is  expressed  by  the  ratio  of  the  heat  equivalent  of  the  work  done  on 
the  piston  per  unit  of  time,  to  the  heat  supplied.     The  degree  of  per- 
fection realized  is  ascertained  by  comparing   the  performance  of  the 
real  engine  with  that  of  an  ideally  perfect  engine,  working  between  the 
same  temperature  limits.     The  theoretical  limit  of  perfection  is  that 
defined   by  the  Carnot    cycle,  the   efficiency  of  which   is  represented 
by  the  equation 


in  which 

Tl  =  the  highest  absolute  temperature  of  the  working  fluid. 
T2  =  the  lowest  absolute  temperature  of  the  working  fluid. 

The  upper  limit  of  temperature  is  that  corresponding  to  boiler  pressure, 
and  the  lower  limit  to  that  of  the  exhaust  steam.  Evidently  the 
greater  the  temperature  range  the  more  nearly  does  the  ideal  efficiency 
approach  unity,  but  with  the  present  limits  of  temperature  used  in 
steam  engines,  it  cannot  exceed  about  35  per  cent. 

The  nearest  approach  of  any  actual  engine  to  the  Carnot  cycle  is 
accomplished  by  the  Nordburg  system  of  progressive  feed  heating, 

267 


268  STEAM  POWER  PLANT  ENGINEERING 

in  which  the  feed  water  is  successively  heated  from  the  receivers  inter- 
mediate between  each  pair  of  cylinders.  (Engineering  News,  May  4: 
1899,  p.  283.)  Table  36  gives  the  Carnot  efficiencies  of  condensing  and 
non-condensing  engines  for  ordinary  ranges  of  steam  pressures. 

The  Carnot  cycle  is  theoretically  impossible  for  an  engine  using 
superheated  steam  at  constant  pressure,  and,  in  general,  it  is  not  very 
closely  simulated  by  engines  using  saturated  steam.  It  is,  therefore, 
more  instructive  to  select  an  ideal  cycle  which 
more  nearly  represents  the  performance  of  the 
actual  engine.  The  diagram  representing  the 
operation  of  this  perfect  engine  is  shown  in 
Fig.  142a,  and  is  called  the  non-conducting  or 
Rankine  cycle,  ab  represents  the  admission 
of  dry  steam  from  the  boiler  at  pressure  p^, 
be  is  an  adiabatic  expansion  to  exhaust  pressure 
p2;  cd  represents  the  exhaust,  and  da  is  an  adia- 
batic compression  to  the  initial  pressure. 

The  heat  necessary  to  raise  the  feed  water  from  the  temperature  of 
exhaust,  or  ideal  feed  water  temperature,  to  the  temperature  in  the  boiler 
and  evaporate  it  into  dry  steam  is 

#1  =  r,  +  q,  -  q2,  (66) 

in  which 

Hl  =  quantity  of  heat  supplied  to  the  cylinder  per  pound  of  steam. 
rt  =  heat  of  vaporization  at  pressure  pr 
ql  =  heat  of  the  liquid  at  pressure  pv 
q2  =  heat  of  the  liquid  at  pressure  p2. 

The  heat,  H2,  exhausted  from  the  cylinder  and  which  must  be  with- 
drawn when  it  is  condensed  is 

#2  =  x2r2)  (66a) 

in  which 

x2  =  quality  of  the  steam  at  pressure  p2. 
r2  =  heat  of  vaporization  at  pressure  p2. 

x2  may  be  calculated  by  the  aid  of  equation 

(66b) 


STEAM  ENGINES 


269 


270 


STEAM  POWER  PLANT  ENGINEERING 


STEAM  ENGINES  271 

in  which 

T2  =  absolute  temperature  of  steam  at  pressure  p2. 
Tl  =  absolute  temperature  of  steam  at  pressure  pr 

61  =  entropy  of  the  liquid  at  pressure  pt. 

62  =  entropy  of  the  liquid  at  pressure  p2. 

Other  notations  as  above. 

The  heat  changed  into  work  per  pound  of  steam  is 

Hl-H2  =  rl+ql-q2-  X2r2,  (67) 

and  the  efficiency,  Er,  of  the  cycle  is 


Er  =     V2  =    l        i  2-  (67a) 

Hi  ri   +  01   ~  & 

The  steam  consumption  W,  or  water  rate,  Ibs.  per  h.p.  hr.  of  the  perfect 
engine,  may  be  expressed 


W  =  --  (67b) 

*1  +  $1  -   ?2 

If  the  steam  entering  the  cylinder  is  wet  and  of  quality  xlf  substitute 
x^i  in  above  equations  for  rt. 

If  the  steam  is  superheated  at  admission  but  becomes  moist  at  the 
lower  pressures,  which  is  the  usual  case,  the  efficiency  may  be  expressed 


1        l222> 

'l    +   01    +   CJ*   ~   ?2 

in  which 

ct  =  mean  specific  heat  of  the  superheated  steam  at  pressure  pt. 
ts  —  degree  of  superheat  or  difference  in  temperature  between  the 
superheated  and  saturated  steam  at  pressure  pr 

x2  may  be  calculated  by  the  aid  of  equation 


i         I      i     ft  ~   2     I     a  f£.'7A\ 

*/      ~T       ~T        l  =  ~T~       2'  (67d) 

in  which 

t  and  T  =  thermometric  and  absolute  temperatures  of  the  superheated 

steam. 
c  =  true  specific  heat  of  superheated  steam  at  temperature  t. 

Other  notations  as  above. 


272 


STEAM  POWER  PLANT  ENGINEERING 


STEAM  ENGINES  273 

For  many  purposes  equation  (67d)  may  be  expressed 

e,  log.  £  +  £  +  0,  -  ^  +  «,.  (67e) 

•L  s  *  i  *  2 

For  highly  superheated  steam  in  which  the  steam  is  still  superheated 
at  exhaust 

E   =  rl  +  qi  +  Clt.  -r2-q2-  cj.'t  (67f 

r,  +  ?i  +  ct«,  -  & 
in  which 

c2  =  mean  specific  heat  of  the  superheated  steam  at  exhaust. 

ts'  =  degree  of  superheat  at  exhaust. 

tsr  may  be  calculated  by  the  aid  of  equation 


Problems  connected  with  the  Rankine  cycle  may  be  conveniently 
solved  by  temperature-entropy  tables  to  be  found  in  connection  with 
the  usual  steam  tables  or  by  the  Mollier  diagram  as  described  in 
Appendix  H. 

158.  Thermal  Efficiency  of  the  Actual  Engine.  —  In  calculating  the 
thermal  efficiency  of  the  real  engine  the  heat  supplied  is  reckoned  above 
the  sensible  heat  of  the  exhaust,  thus, 

„  _  Heat  converted  into  useful  work 
Heat  supplied 

(69) 


B.T.U.  supplied  per  I.H.P.  per  minute 
42.42 


w 
in  which 


(70) 


w  =  the  weight  of  steam  supplied  to  the  engine  per  indicated  horse 
power  per  minute,  or  per  brake  horse  power  per  minute,  de- 
pending upon  whether  the  efficiency  is  to  be  referred  to  the 
indicated  or  to  the  brake  horse  power  of  the  engine. 

Other  notations  as  in  (66)  and  (67). 

The  figure  obtained  by  dividing  the  efficiency  of  the  real  engine  by 
that  of  the  ideal  engine  is  called  the  efficiency  ratio,  and  is  a  measure  of 
the  extent  to  which  the  theoretical  possibilities  are  realized. 


274 


STEAM  POWER  PLANT  ENGINEERING 


The  efficiency  ratio  is  calculated  on  the  basis  of  the  indicated  horse 
power  or  the  developed  horse  power: 


Eff.  Ratio  =  -=-' 
h>T 


42.42 


w  (H,  -  H2) 


(71) 


This  ratio  expressed  in  different  terms  has  been  referred  to  as  "  Poten- 
tial Efficiency"  by  C.  V.  Kerr  (Trans.  A.S.M.E.,  25-920),  and  as 
"  Cylinder  Efficiency  "  by  Professor  Reeve. 

The  commercial  economy  of  an  engine  is  measured  by  the  cost  of 
producing  power,  and  does  not  necessarily  depend  upon  its  thermal 
efficiency.  The  performances  of  steam  engines  are  frequently  stated 
in  terms  of  (a)  pounds  of  steam  utilized  per  horse-power  hour,  (b) 
pounds  of  coal  per  horse-power  hour,  .(c)  cost  in  cents  per  horse-power 
hour,  (d)  B?T.U.  per  horse-power  hour.  From  a  commercial  stand- 


TABLE   36. 
STEAM-ENGINE    EFFICIENCIES. 

(Saturated  Steam.) 


Non-Condensing;  Back  Pressure  14.7 

Condensing;  Back  Pressure  1  Pound 

Absolute. 

Absolute. 

Gauge 

Press. 

Ratio 

Ratio 

Carnot 

Rankine 

Actual  * 

Carnot 

Rankine 

Actual 

Cycle. 

Cycle  (a). 

(b). 

-• 

Cycle. 

Cycle  (a). 

(b). 

!?. 

a 

a 

25 

7.5 

7.3 

5.5 

76°0 

22.6 

21.0 

11.6 

55°0 

50 

11.2 

10.7 

8.5 

80 

25.7 

23.5 

13.5 

60 

75 

13.7 

13.0 

10.4 

80 

27.8 

25.3 

15.9 

61 

100 

15.7 

14.8 

12.0 

81 

29.5 

26.7 

20.2 

76 

125 

17.3 

16.3 

13.5 

83 

•      30.8 

27.8 

20.3 

74 

150 

18.7 

17.5 

14.3 

82 

32.0 

28.8 

21.6 

75 

175 

19.8 

18.5 

14.8 

80 

32.9 

29.6 

21.9 

74 

200 

20.8 

19.3 

15.2 

79 

33.7 

30.2 

22.6 

75 

225 

21.6 

19.9 

15.5 

78 

34.5 

30.6 

22.6 

74 

250 

22.4 

20.5 

35.1 

31.0 

275 

23.0 

21.0 

35.6 

31.3 

300 

23.6 

21.4 

36.0 

31.5 

*  Best  recorded  performance  of  the  actual  engine,  1907. 

point  the  cost  of  producing  power  is  the  most  important  basis  of  com- 
parison, but  the  latter  expression  is  most  satisfactory  for  scientific 
purposes,  since  it  gives  a  basis  of  comparing  the  performances  of  all 
types  of  heat  engines. 


STEAM  ENGINES 


275 


159.  Mechanical  Efficiency.  —  The  power  of  an  engine  may  be 
expressed  in  terms  of  indicated  horse  power,  brake  horse  power, 
or  pump  horse  power,  according  to  the  class  of  engine.  The  ratio 
of  the  brake  to  the  indicated  power  is  the  mechanical  efficiency 
of  the  engine,  the  ratio  of  the  electric  horse  power  to  the  indicated 
power  is  the  mechanical  efficiency  of  the  engine  and  generator  com- 
bined,, and  the  ratio  of  the  pump  horse  power  to  the  indicated  power 
of  the  steam  engine  is  the  mechanical  efficiency  of  the  engine  and  pump 


•Jo 
90 
85 
80 
75 
70 
G5 

t  J 

r^- 

— 

^^ 

—  —  - 

_      - 

4 

^ 

x 

^ 

^0 

o 

^^_ 

\ 

• 

c/ 

4 

^ 

^ 

xx"" 

/ 

/ 

•^ 

^ 

/ 

/ 

/ 

Mechanical 
Efficiencies  of 
75  K.W.Geuerating  Set 
Engine,  Simple  High  Speed 
Non  Condensing 

E 

7 

/ 

GO 
55 
50 
45 

/ 

/° 

i 

/ 

°i 

/ 

)       10       20      30      40       50       60       70       80       90      100     110     120     130     140     15 
Per  Cent  of  Rated  Load 

FIG.  143. 


combined.     The  percentage  of  work  lost  in  friction  is  therefore  the 
difference  between  100  per  cent  and  the  mechanical  efficiency. 

Table  37  shows  the  mechanical  efficiencies  for  several  types  of  en- 
gines, and  Fig.  143  the  combined  efficiency  of  a  direct-connected  high- 
speed engine  and  generator.  (See  Engine  Friction,  par.  167.) 

Mechanical  Efficiency:  Peabody,  Thermodynamics,  p.  430;  Spangler,  Steam 
Engineering,  p.  205;  Ripper,  Steam  Engine,  p.  275;  Ewing,  Steam  Engine,  p.  186. 

The  following  numerical  example  will  illustrate  the  calculation  of 
the  various  efficiencies  mentioned: 

A  simple  high-speed  engine  uses  30  pounds  of  steam  per  I.H.P. 
hour;  initial  pressure  100  pounds  per  square  inch,  gauge;  exhaust 
pressure,  atmospheric;  I.H.P.,  120;  D.H.P.,  102.  Steam  assumed  to 
be  dry  and  saturated  at  throttle.  Required:  (1)  the  actual  thermal 
efficiency;  (2)  the  efficiency  of  the  Carnot  cycle;  (3)  the  efficiency  of 
the  Rankine  cycle;  (4)  the  efficiency  ratio;  (5)  the  mechanical 
efficiency. 


276 


STEAM  POWER  PLANT  ENGINEERING 


TABLE  37. 
MECHANICAL    EFFICIENCIES    OF   ENGINES. 


Kind  of  Engine. 

Horse  Power. 

Efficiency  at  Full 
Load. 

Simple  : 
1     High-speed,  non-condensing 

150 

95  5  ^ 

2    High-speed,  condensing 

170 

96 

3    Low-speed,  non-condensing  .       .  . 

275 

94 

4.  Low-speed,  condensing  

Compound  : 
5    High-speed   non-condensing 

150 

94  « 

6    High-speed,  condensing       .... 

160 

98 

7    Low-speed,  non-condensing     

900 

95 

8.   Low-speed,  condensing  

1000 

95 

Triple:     (combined    efficiency    of    engine    and 
pump) 
9    Pumping  engine 

865 

97  4 

Quadruple:  (combined  efficiency  of  engine  and 
pump) 
10    Pumping  engine  

712 

93 

1.  Buffalo  Simple  engine,  12  X  12,  Elec.  World,  Sept.,  1904,  p.  147. 

2.  Reeves  Simple  engine,  15  X  14,  Elec.  World,  Oct.  1,  1904,  p.  587. 

3.  24  X  48  Hamilton  Corliss  at  Armour  Inst.  of  Tech.,  1898. 
4. 

5.  Reeves  Compound;  Eng.  Rec.,  July  1,  1905,  p.  24. 

6.  Reeves  Compound;  Eng.  Rec. 

7.  21,  41  X  30  Cross  Compound  Ball  &  Wood,  West  Albany  Station,  N.Y.C.  &  H.R.R, 

8.  20,  40  X  42  Rice  &  Sargent;  A.S.M.E.,  29-1276. 

9.  Allis  Pumping  Engine;  Power,  May,  1906,  p.  299. 

10.  Nordburg  Pumping  Engine;  Eng.  News,  May  4,  1899,  p.  280. 

(1)  The  actual  thermal  efficiency  is 

42.42 


E 


W 


ql  -  q2) 
42.42 


30/60  (879.6  +  308.9  -  180.3) 
=  0.084. 

That  is,  only  8.45  per  cent  of  the  heat  supplied  to  the  cylinder  above 
the  temperature  of  the  exhaust  steam  is  converted  into  work.  The 
assumption  that  the  exhaust  steam  may  be  used  to  heat  the  feed  water 
to  its  own  temperature  justifies  reckoning  the  heat  supplied  above  the 
exhaust  temperature. 

(2)  The  efficiency  of  the  Carnot  cycle  is 


798.8-673 
798.8 


=  0.157, 


STEAM  ENGINES 


277 


which  means  that  if  the  engine  were  a  perfect  one  employing  the  Car- 
not  cycle  between  the  same  extremes  of  temperature  as  the  actual 
engine,  15.7  per  cent  of  the  energy  supplied  would  be  converted  into 
useful  work. 

(3)  On  the  basis  of  the  Rankine  cycle  the  ideal  efficiency  becomes 

7?  Xiri   ~J~   ?1    ~   X2T2   ~~   <?2 


879.6  +  308.9  -  0.885  X  969.7  -  180.3 


=  0.149, 


879.6  +  308.9  -  180.3 


the  numerator  representing  the  heat  converted  into  work  per  pound  of 
steam  and  the  denominator  the  heat  supplied.     Thus  if  the  engine  be 


4        6        8       10      12       14      16      18       20      22      24      26       28      30      32      34 
Thermal  Efficiency,  Per  Cent 

FIG.  144. 


assumed  to  have  a  non-conducting  cylinder  and  work  on  the  Rankine 
cycle,  it  would  be  capable  of  utilizing  14.9  per  cent  of  the  heat  sup- 
plied. 

(4)  The  efficiency  ratio  is  expressed: 

Efficiency  ratio  =  f*  =  |^|  =  56.7  per  cent, 
Ej        14.9 


278  STEAM  POWER  PLANT  ENGINEERING 

which  indicates  the  degree  of  perfection  of  the  engine  or  the  extent  to 
which  it  realizes  the  maximum  efficiency  theoretically  possible.  The 
weight  of  steam  which  would  have  to  be  consumed  per  horse  power 
per  hour  to  actually  attain  the.  Rankine  efficiency  would  be 

30  X  ^^  =  30  X  0.567  =  17  pounds. 

(5)  The  mechanical  efficiency  =  ^Jf-JL'   =  ^=0.85  =  85  per  cent. 

I.  ii.  Jr.         120 

The  relation  between  the  actual  and  the  theoretical  efficiencies 
based  upon  the  Rankine  cycle,  and  upon  the  best  recorded  perform- 
ances of  modern  engines,  using  saturated  steam,  is  shown  graphically  in 
Fig.  144.  (Also  see  Tables  39  and  40.)  The  theoretical  curves  are 
calculated  upon  the  assumption  of  complete  expansion,  the  back 
pressure  being  14.7  pounds  gauge  for  non-condensing  and  one  pound 
absolute  for  condensing  engines. 

The  highest  recorded  (1907)  efficiency  ratio  for  a  steam  engine 
using  saturated  steam  is  83  per  cent  non-condensing  and  76  per  cent 
condensing  (see  Table  39),  and  78.5  per  cent  for  a  condensing  engine 
using  superheated  steam  (Table  43). 

160.  Heat  Losses  in  the  Steam  Engine.  —  The  principal  losses  which 
tend  to  lower  the  efficiency  of  the  steam  engine  are  due  to 

(a)  Presence  of  moisture  in  the  steam  at  admission. 

(6)  Leakage  past  valves  and  piston. 

(c)  Cylinder  condensation. 

(d)  Clearance  volume. 

(e)  Incomplete  expansion. 
(/)  Wire  drawing. 

(g)  Friction. 
(h)  Radiation. 

161.  Moisture.  —  The  presence   of   moisture  in   the  steam   pipe  is 
due  to  condensation  caused  by  radiation  or  to  priming  at  the  boiler. 
Unless  removed  by  some  separating  device  between  boiler  and  engine 
the  amount  of  moisture  entering  the  cylinder  may  be  from  1  to  5  per 
cent  of  the  total  weight  of  steam,  and  the  work  done  per  pound  of 
fluid  is   correspondingly   reduced.     This   loss  should  not   be   charged 
against  the  engine,  however,  and  its  performance  should  be  reckoned 
on  the  dry  steam  basis.     Experiments  reported  by  Professor  R.  C.  Car- 
penter (Trans.  A.S.M.E.,  15  —  438)  in  which  water  in  varying  quantities 
was  introduced  into  the  steam  pipe,  causing  the  quality  of  the  steam 
to  range  from  99  per  cent  to  57  per  cent,  showed  that  the  consumption 


STEAM  ENGINES  279 

of  dry  steam  per  I.H.P.  hour  was  practically  constant,  the  water  acting 
as  an  inert  quantity.  An  efficient  separator  will  remove  practically 
all  the  entrained  water. 

Influence  of  Moisture  on  Steam  Economy:  Trans.  A.S.M.E.,  18-699;  Benjamin, 
Heat  and  Steam;  Perry,  Steam  Engine,  p.  353;  Rankine,  Steam  Engine,  p.  407; 
Ripper,  Steam  Engine,  p.  33;  Ewing,  The  Steam  Engine,  p.  139. 

162.  Leakage  of  Steam.  —  The  loss  due  to  leakage   is  a  variable 
factor  depending  upon  the  design  and  condition  of  the  engine,  and  is 
greater  with  saturated  than  with  superheated  steam.     The  usual  method 
of  measuring  leakage  past  the  valves  and  piston  while  the  engine  is  at 
rest  is  likely  to  give  erroneous  results  as  demonstrated  by  Callender 
and  Nicolson  (Peabody,  "Thermodynamics,"  p.  351)  in  tests  made  on  a 
high-speed  automatic  balanced  valve  engine  and  on  a  quadruple  expan- 
sion engine  with  plain  unbalanced  slide  valves.     With  the  engines  at 
rest  they  found  that  the  leakage  past  valves  and  piston  was  insignificant, 
but  when  in  operation  the  leakage  from  the  steam  chest  into  the  exhaust 
was  very  considerable  indeed.     It  was  thought  that  a  large  proportion 
of  the  leakage  was  probably  in  the  form  of  water  formed  by  condensa- 
tion of  steam  on  the  seat  uncovered  by  the  valve. 

According  to  the  report  of  the  Steam  Engine  Research  Committee 
(Eng.  Lond.,  March  24,  1905,  p.  298),  leakage  through  a  plain  slide 
valve  is  independent  of  the  speed  of  the  sliding  surfaces,  and  directly 
proportional  to  the  difference  in  pressure  on  the  two  sides;  with  well- 
fitted  valves  the  leakage  is  never  less  than  4  per  cent  of  the  volume  of 
steam  entering  the  cylinders,  and  is  often  greater  than  20  per  cent. 

Peabody,  Thermodynamics,  p.  350;  Eng.  Rec.,  May  22,  1897,  p.  529;  Barrus, 
Engine  Tests,  p.  251. 

163.  Cylinder    Condensation.  —  A    large    percentage   of   the   steam 
admitted  to  the  cylinder  is  condensed,  due  to  the  absorption  of  heat 
by  the  relatively  cool  cylinder  walls.     Condensation  continues  during 
expansion  until  the  temperature  of  the  steam  falls  below  that  of  the 
metal,  when  the  process  is  reversed  and  a  part  of  the  moisture  is  re- 
evaporated.     Unless  the  cylinder  is  one  of  a  compound  series,  the  heat 
absorbed  from  the  cylinder  walls  during  exhaust  does  no  useful  work 
and  is  lost.     The  condensation  up  to  the  point  of  cut-off  (initial  con- 
densation) may  amount  to  from  15  to  30  per  cent,  and  is  often  as  high 
as  50  per  cent  of  the  total  weight  admitted  to  the  cylinder.     The 
initial  condensation  becomes  greater  as  the  difference  between  initial 
and  exhaust  pressures  is  increased,  and  diminishes  as  the  speed  of  the 
engine  increases.     Cylinder  condensation  and  leakage  are  ordinarily 


280 


STEAM  POWER  PLANT  ENGINEERING 


classified  together,  as  there  is  no  way  of  separating  them  accurately. 
They  represent  that  part  of  the  feed  water  which  is  not  accounted  for 
by  the  indicator  diagram. 

Tests  of  20  simple  high-speed  engines  by  G.  H.  Barrus,  Fig.  145, 
show  some  results  obtained  for  various  percentages  of  cut-off.  Also 
see  table  compiled  by  C.  H.  Peabody,  "  Thermodynamics  of  the  Steam 


50 


40 


10 


o 

\ 

\ 

Condensation.and  Leakage 
for 
Simple  Engines 
using 
Saturated  Steam* 

s 

o 

\ 

S 

\o 

N, 

0 

V 

O 

o 

^X 

n 

\° 

D 
O 

^ 

< 

>r 

^ 

5     r 

^ 

^—-  ^ 

Cr? 

O 

E 

igine 

Tests 

Ba" 

•us,  p. 

254. 

0  5  10  15  20  25'  30  35  40 

Percentage  of  Cut  Off 

FIG.  145. 

Engine,"  p.  336,  showing  analysis  of  the  heat  interchanges  for  a  number 
of  different  types  of  steam  engine. 

The  various  heat  losses,  including  cylinder  condensation  and  leakage, 
are  best  determined  by  transferring  the  indicator  diagram  to  the  tem- 
perature entropy  or  6<j>  chart.  (See  Appendix  C.)  This  is  useful  for 
certain  scientific  investigations,  but  is  unnecessary  for  commercial  tests. 

Cylinder  Condensation :  Trans.  A.S.M.E.,  I,  184,  III,  215,  IV,  88,  VII,  375, 
XVIII,  950;  Cotterill,  Steam  Engine,  p.  331;  Spangler,  Steam  Engineering,  p.  228; 
Thurston,  Manual  of  the  Steam  Engine,  I,  271,  488,  585;  Heck,  Steam  Engine, 
pp.  109,  113,  119;  Ewing,  Steam  Engine,  p.  148;  Hutton,  Heat  Engines,  p.  319; 
Peabody,  Thermodynamics,  pp.  241,  412;  Reeve,  Thermodynamics,  p.  198;  Wood, 
Thermodynamics,  p.  212;  Perry,  Steam  Engine,  p.  78;  Ripper,  Steam  Engine,  p.  25. 

Initial  Condensation:  Cotterill,  Steam  Engine,  p.  274;  Golding,  6<j>  Diagram, 
p.  63;  Marks,  Steam  Engine,  p.  195;  Peabody,  Thermodynamics,  p.  359;  Popplewell, 
Heat  Engine,  pp.  323,  351;  Reeve,  Thermodynamics,  pp.  156,  221;  Ripper,  Steam 
Engine,  pp.  Ill,  168. 


STEAM  ENGINES  281 

Condensation  during  Expansion:  Trans.  A.S.M.E.,  III,  286;  Hutton,  Heat 
Engines,  pp.  223,  286;  Pupin,  Thermodynamics,  p.  88;  Rankine,  Steam  Engine, 
p.  385;  Reeves,  Thermodynamics,  p.  221. 

Entropy  :  Power,  Jan.  21,  1908;  Baynes,  Thermodynamics,  p.  94;  Benjamin, 
Heat  and  Steam,  p.  37;  Berry,  Temperature-Entropy  Diagram;  Boulvin,  Entropy 
Diagram;  Ewing,  Steam  Engine,  p.  103;  Golding,  0<j>  Diagram;  Hutton,  Heat 
Engines,  p.  276;  Peabody,  Thermodynamics,  p.  97  ;  Reeve,  Thermodynamics,  p.  39; 
Swinburne,  Entropy;  Wood,  Thermodynamics,  p.  136:  Heck,  Steam  Engine, 
Chap.  VI. 

164.  Clearance  Volume.  —  The  portion  of  the  cylinder  volume  not 
swept  through  by  the  piston  but  which  is  nevertheless  filled  with 
steam  when  admission  occurs  is  called  the  clearance  volume.  It  is  the 
space  between  the  end  of  the  piston  when  on  dead  center  and  the 
inside  of  the  valves  covering  the  ports.  It  varies  from  about  1  per  cent 
of  the  piston  displacement  in  very  large  engines  with  short  steam 
passages  to  10  per  cent  or  more  in  small  high-speed  engines.  When 
the  steam  retained  in  the  clearance  space  is  compressed  to  the  initial 
pressure  and  expansion  is  carried  down  to  the  back  pressure,  the  clear- 
ance has  little  effect  upon  the  economy  of  the  engine,  but  since  expan- 
sion and  compression  are  seldom  complete  in  actual  practice,  the  loss 
may  be  considerable.  (Ripper,  "  Steam  Engine,"  p.  103.)  The  shorter 
the  cut-off  the  greater  will  be  the  ratio  of  the  weight  of  cushion  steam 
to  that  of  the  steam  supplied  and  hence  the  greater  the  relative  loss. 
In  large  slow-speed  engines  the  loss  may  be  insignificant  if  the  clearance 
volumes  are  small,  while  in  small  high-speed  engines  it  may  be  con- 
siderable. 

The  ratio  of  expansion  is  decreased  by  clearance;  for  example,  an 
engine  cutting  off  at  one-fifth,  neglecting  clearance  has  an  apparent 
ratio  of  expansion  of  5,  but  if  the  clearance  volume  is  10  per  cent  the 
actual  ratio  is  only  3.66.  One  of  the  few  recorded  tests  relative  to  the 
influence  of  clearance  on  the  economy  of  a  high-speed  engine  was  con- 
ducted on  a  14  x  15  Allfree  engine.  (Power,  May,  1901.)  With  a 
clearance  volume  of  2.2  per  cent,  initial  pressure  105  pounds  gauge, 
and  172  r.p.m.,  the  best  performance  was  23.7  pounds  of  dry  steam 
per  I.H.P.  hour.  With  the  same  steam  pressure  and  speed,  but 
with  clearance  volume  increased  to  6  per  cent  by  the  use  of 
a  shorter  piston,  the  best  performance  was  28.3  pounds  per  I.H.P. 
hour.  In  both  cases  the  compression  was  carried  up  to  admission 
pressure. 

Loss  by  Clearance  in  Steam  Engines  :  Trans.  A.S.M.E.,  XVIII,  176.  Clearance  in 
Compound  Engines:  ibid.,  I,  173.  Clearance  in  Multi-cylinder  Engines:  ibid.,  XI,  151. 
Effect  of  Clearance:  Ewing,  Steam  Engine,  p.  145;  Hutton,  Heat  Engines,  p.  334; 


282 


STEAM   POWER  PLANT  ENGINEERING 


Peabody,  Thermodynamics,  p.  407;  Popplewell,  Heat  Engines,  p.  332;  Reeve,  Ther- 
modynamics, pp.  223,  256 ;  Ripper,  Steam  Engine,  pp.  63,  101 ;  Spangler,  Steam 
Engineering,  p.  114;  Wood,  Thermodynamics,  p.  197. 

165.  Loss  Due  to  Incomplete  Expansion  and  Compression.  —  Theo- 
retically the  loss  due  to  incomplete  expansion  is  considerable.  For 
example,  the  theoretical  steam  consumption  of  a  perfect  engine  (Ran- 
kine  cycle)  expanding  from  120  pounds  absolute  to  a  condenser  pres- 
sure of  2  pounds  absolute  is  9.6  pounds  per  horse-power  hour.  If  the 
expansion  were  carried  to  only  5  pounds  absolute,  the  exhaust  pressure 
remaining  the  same,  the  steam  consumption  would  be  increased  to 
11.8  pounds  per  horse-power  hour,  a  difference  of  22  per  cent  for  an 
increase  in  terminal  pressure  of  only  3  pounds  per  square  inch.  The 
theoretical  water  rates  for  various  terminal  pressures  are  given  below. 


Terminal  Pressure, 
Pounds  per  Square 
Inch  Absolute. 

Steam  Consumption 
of  Perfect  Engine. 

Terminal  Pressure, 
Pounds  per  Square 
Inch  Absolute. 

Steam  Consumption 
of  Perfect  Engine. 

1 
1.5 
2 
2.5 

8.5 
9.1 
9.6 
10 

3 
4 
5 
6 

10.4 
11.1 
11.8 
12.3 

In  actual  engines  expansion  is  seldom  complete,  since  it  would 
necessitate  increased  bulk  and  weight  of  engine,  and  the  work  done 
by  the  steam  in  the  last  stages  would  not  compensate  for  the  increased 
cost. 

In  single-cylinder  engines  maximum  economy  is  effected  when  the 
terminal  pressure  is  considerably  above  that  of  the  exhaust,  since  the 
gain  due  to  complete  expansion  is  more  than  offset  by  the  increased 
cylinder  condensation.  This  is  true  to  a  certain  extent  in  all  engines 
irrespective  of  the  number  of  cylinders.  Tests  by  G.H.  Barrus  ("Engine 
Tests,"  1900)  to  determine  the  terminal  pressures  effecting  maximum 
economy  for  various  types  of  engine  gave  results  as  follows : 


Terminal  Pressure, 
Pounds  Absolute. 


Simple  slide-valve  engines,  non-condensing  . 

Simple  slide-valve  engines,  condensing 

Simple  Corliss  engines,  non-condensing 

Simple  Corliss  engines,  condensing 

Compound  engines,  non-condensing 

Compound  engines,  condensing 


30  to  40 
25  to  30 
20  to  25 
15  to  18 
18  to  22 
3  to  5 


STEAM  ENGINES 


283 


In  high-speed  engines  a  certain  amount  of  compression  is  desirable 
for  its  cushioning  effect ;  outside  of  this  mechanical  feature  compression 
may  or  may  not  be  of  benefit  to  the  engine  as  will  be  explained  later. 
Zuener  in  his  treatise  on  theoretical  thermodynamics  proves  deduc- 
tively that  in  an  engine  with  a  large  clearance  volume  the  loss  due  to 
clearance  is  completely  eliminated  if  the  compression  is" carried  up  to 
admission  pressure,  a  conclusion  which  tests  by  Jacobus,  Carpenter, 
and  others  fail  to  confirm.  A  series  of  tests  by  Professor  Jacobus 
(Trans.  A.S.M.E.,  15-918)  on  a  10x11  high-speed  automatic  engine 
at  Stevens  Institute  show  decreasing  economy  with  increase  of  com- 
pression, the  initial  pressure,  cut-off,  and  release  remaining  constant. 
The  results  were  as  follows: 


Proportion   of  initial   pressure  up   to  which   the 

steam  is  compressed 

Steam,  pounds  per  I.H.P.  hour 


34.8 


36.7 


Full 
38 


Tests  by  Carpenter  (Trans.  A.S.M.E.,  16-957)  on  the  high-pressure 
cylinders  of  the  Corliss  engine  at  Sibley  College  gave: 


Compression,  per  cent.  .  .  . 

11.4 

25 

35.2 

Brake  horse  power  

30 

29 

26 

Steam  pounds  per  B  H  P 

hour 

33 

33  3 

34 

6.8 
6.7> 
6.6 
6.5 
6.4 
6.3 
6.2 
6.1 
6.0 
5.9 
5.8' 
5.7 
5.6 
5,5 

tial  Pressure 
3  Lb.  Gauge 

10 

^^ 

\ 

^^ 

js^ 

^ 

^ 

^^ 

s^& 

<^& 

v# 

\, 

*j£ 

\ 

^ 

^**^  > 

1 

^^ 

^^X. 

^ 

• 

^^ 

^^ 

Influence  of  Back 
Pressure  on  the  Economy  of 
an  8  x  10  Automatic  High  Speed 
Non  Condensing  Engine 

\^ 

^^ 

? 

6  8  10  12  14 

Back  Pressure.Lb.  Per  Sq,In.Gauge 

FIG.  145a. 


16 


18 


20 


Opposed  to  these  figures  are  tests  which  show  an  improvement  in 
economy  when  compression  is  increased. 

Fig.   145a  shows  the  influence  of  increasing  back  pressure  on  the 


284  STEAM  POWER  PLANT  ENGINEERING 

economy  of  an  8  x  10  automatic  high-speed  engine  at  Armour  Institute 
of  Technology. 

Cut-off. —  Best  for  Different  Pressures:  Trans.  A.S.M.E.,  4-89;  In  Compound 
Engines:  ibid.,  p.  549;  Most  Economical  Point  of:  ibid.,  8-486;  Hutton,  Heat 
Engines,  p.  232 ;  JPeabody,  Thermodynamics,  p.  210;  Spangler,  Steam  Engineering, 
p.  109;  Wood,  Thermodynamics,  pp.  200,  433;  Klein,  High-Speed  Engines,  p.  7; 
Ewing,  Steam  Engine,  p.  84. 

Ratio  of  Expansion:  Trans.  A.S.M.E.,  2-19,  128,  10-576,  11-166;  Ewing,  Steam 
Engine,  pp.  47, 159;  Wood,  Thermodynamics,  pp.  154, 172, 197,  295;  Rankine,  Steam 
Engine,  pp.  378,  553;  Reeve,  Thermodynamics,  p.  228;  Thurston,  Manual  of  the 
Steam  Engine,  1-271,  725,  2-14. 

Incomplete  Expansion:  Boulvin,  Entropy  Diagram,  p.  28;  Cotterill,  Steam  Engine, 
p.  240;  Perry,  Steam  Engine,  p.  364;  Popplewell,  Heat  Engines,  p.  332;  Peabody, 
Thermodynamics,  p.  238;  Reeve,  Thermodynamics,  p.  222;  Heck,  Steam  Engine, 
p.  78. 

Compression.  —  Efficiency  of  Compression  in  Steam  Engine  :  Engr.,  Lond.,  Nov.  3, 
1905,  p.  434;  Compression  as  a  Factor  in  Steam  Engine  Economy:  Trans.  A.S.M.E., 
14-1067;  Effect  of  Compression  on  Water  Consumption:  ibid.,  15-815;  Engine  Com- 
pression: ibid.,  7-708;  In  High-Speed  Engines:  ibid.,  7-202;  In  Steam  Cylinder:  ibid., 
2-341. 

Back  Pressure. —  Back  Pressure  as  Modifying  Economy:  Trans.  A.S.M.E.,  18-283; 
On  Valves :  ibid.,  3-150;  General ;  Ewing,  Steam  Engine,  pp.  84,  145;  Klein,  High- 
Speed  Engines,  p.  11;  Perry,  Steam  Engine,  p.  75;  Reeve,  Thermodynamics,  p.  223; 
Ripper,  Steam  Engine,  p.  53. 

166.  Loss  due  to  Wire  Drawing.  —  Wire   drawing,  or  the  drop  in 
pressure  due  to  the  resistances  of  the  ports  and  passages,  has  the  effect 
of  reducing  the  output  of  the  engine  to  some  extent,  since  the  pressure 
within  the  cylinder  is  less  than  that  at  the  throttle  during  admission 
and  greater  than  discharge  pressure  at  exhaust.     The  steam  may  be 
dried  to  a  small  extent  during  admission.     In  single-valve  engines  the 
effects  of  wire  drawing  are  decidedly  marked  and  the  true  points  of 
cut-off  and  release  are  sometimes  difficult  to  locate  on  the  indicator 
card.     In  engines  of  the  Corliss  or  gridiron-valve  type  the  effects  are 
hardly  noticeable. 

Wire  Drawing:  Trans.  A.S.M.E.,  2-344,  1-174;  Ewing,  Steam  Engine,  pp,  95, 
143,  207;  Boulvin,  Entropy  Diagram,  p.  56;  Popplewell,  Heat  Engine,  p.  320; 
Rankine,  Steam  Engine,  p.  413;  Reeve,  Thermodynamics,  pp.  105,  221;  Ripper, 
Steam  Engine,  p.  73;  Wood,  Thermodynamics,  p.  195;  Heck,  Steam  Engine,  pp.  183, 
224,  230. 

167.  Loss  due  to  Friction  of  the  Mechanism.  —  The  difference  between 
the  indicated  horse  power  and  that  actually  developed  is  the  power 
consumed  in  overcoming  friction,  and  varies  from  4  to  20  per  cent  of 
the  indicated  power,  depending  upon  the  type  and  condition  of  the 
engine.     Engine  friction  may  be  divided  into   (1)  initial  or  no-load 


STEAM  ENGINES 


285 


friction  and  (2)  load  friction.     The  stuffing-box  and  piston-ring  friction 
is  practically  independent  of  the  load,  while  that  of  the  guides,  bearings, 


40 


50 


75         100        125        150        175        200        225        250        275        300 
Developed  Harse-Power 


FIG.  146.     Typical  Curves  of  Steam  Engine  Friction. 

and  the  like  increases  with  the  load.  In  Fig.  146,  curve  A  gives  the 
relation  between  the  frictions  for  a  four-slide-valve  horizontal  cross 
compound  engine,  and  B  that  for  a  simple  non-condensing  Corliss. 

TABLE  38. 

DISTRIBUTION  OF  FRICTION  IN  SOME  DIRECT-ACTING  STEAM  ENGINES. 

(Thurston.)* 


Parts  of  Engines  where  Friction 
is  Measured. 

Percentage  of  Total  Engine  Friction. 

S:Js 
•liU 

1=   1  'j 

2  2.0  Z 
*%g£ 

->£  jj 

d  if* 

o  c  a  v 

£« 

11* 

j^s 

§  d  c  o5 

1  1ri 

g  w»3> 

122 

|Il2 

Main  bearings  

47.0 

35.4 

35.0 

41.6 

46.0 

Piston  and  piston  rod  

32.9 

25.0 

21.0 

49.1 

21.8 

Crank  pin  

6.8 

5.1 

13.0 

Crosshead  and  wrist  pin  

5.4 

4.1 

Valve  and  valve  rod  

2.5 

26.4 

22.0 

9.3 

21.0 

Eccentric  strap  

5.4 

4.0 

Link  and  eccentric  

9.0 



Air  pump  

12.0 

100.0 

100.0 

100.0 

100.0 

100.0 

*  "  Friction  and  Lost  Work  in  Machinery,"  p.  13. 


286 


STEAM  POWER  PLANT  ENGINEERING 


(Peabody's  "  Thermodynamics/'  pp.  433  and  437.)  Curve  C  is  plotted 
from  the  tests  of  a  Reeves  vertical  cross^  compound  condensing  engine 
(Engineering  Record,  July  1,  1905,  p.  24),  and  D  from  the  test  of  an  Ames 
simple  high-speed  non-condensing  engine.  (Engineering  Record,  Vol.  27, 
p.  225.)  A  large  number  of  recorded  tests  show  less  friction  at  full  load 
than  at  no  load,  but  this  is  probably  due  to  error  or  to  variations 
in  lubrication.  With  first-class  lubrication  it  is  usually  sufficiently 
accurate  to  assume  the  friction  to  be  constant  and  equal  to  the  initial 
friction  at  zero  load.  The  distribution  of  the  frictional  losses  in  a 
number  of  engines  is  given  in  Table  38. 

Friction  in  Engines  :  Trans.  A.S.M.E.,  8-86,  10-10,  8-108,  9-74,  82,  7-639,  641, 
1-153;  Ewing,  Steam  Engine,  p.  186;  Ripper,  Steam  Engine,  p.  275;  Peabody, 
Thermodynamics,  p.  430;  Perry,  Steam  Engine,  p.  270;  Ripper,  Steam  Engine,  1-540; 
Heck,  Steam  Engine,  316-318. 

The  efficiency  of  the  fluid  in  the  steam  engine  cylinder  may  be  in- 
creased by  (1)  raising  the  boiler  pressure,  (2)  compounding,  (3)  use  of 
reheater-receiver,  (4)  steam  jacketing  the  cylinders,  (5)  increasing  the 
rotative  speed,  (6)  superheating,  (7)  diminishing  the  back  pressure  by 
securing  a  more  perfect  vacuum. 

168.  Effect  of  Increased  Steam  Pressure.  —  A  consideration  of  the 
Rankine  and  Carnot  cycles  indicates  that  theoretically  the  greater  the 
temperature  range  the  greater  will  be  the  efficiency.  (See  Table  36.) 
In  the  actual  engine  the  temperature  range  is  most  readily  increased 
by  raising  the  boiler  pressure,  since  the  limit  of  the  back  pressure  is 
practically  fixed  by  the  cooling  medium  in  the  condenser.  The  theoreti- 
cal gain  resulting  from  increased  pressure  range  is,  however,  very  con- 
siderably affected  by  the  increased  losses  due  to  cylinder  condensation. 

Fig.  147  shows  the  results  of  tests  made  at  the  Armour  Institute  of 
Technology  on  an  8  x  10  automatic  high-speed  piston-valve  engine, 
showing  marked  gain  with  increase  of  initial  pressure  up  to  a  certain 
point  when  the  condensation  losses  became  sufficiently  great  to  neu- 
tralize the  advantage  which  would  otherwise  be  gained. 

The  following  figures  were  obtained  in  tests  of  a  small  Willans  engine, 
non-condensing,  under  different  steam  pressures : 


Initial  Pressure,  Gauge. 

Pounds  Steam  per  I.H.P. 
Hour. 

B.T.U.  per  I.H.P.  per 
Minute. 

36.3 

42.8 

700 

51.0 

36.0 

595 

74.0 

32.6 

544 

85.0 

29.7 

495 

97.0 

26.9 

450 

110.0 

27.8 

465 

122.0 

26.0 

436 

STEAM  ENGINES 


287 


Referring  to  Table  36,  it  may  be  noted  that  both  the  theoretical 
and  the  actual  efficiencies  increase  very  slowly  for  pressures  above  150 
pounds.  Practically,  gain  in  efficiency  due  to  increasing  the  pressure 


Thermal  Efficiency,  Per  Cent 

"  PI  pT  OS  05  05 

i  oo  co  b  MI  to 

\ 

- 

S  fe  fc  &  Ss  ft  6 

Pounds  of  Steam  per  Developed  H.P.-Hr. 

X 

s* 

""  . 

N.  >e 

> 

/ 

^ 

^X 

i 

x 

X 

< 

^ 

0< 

V 

- 

/ 

X 

^ 

^^ 

_ 

• 

^ 

""•*• 

75     80     85      90     95     100     105     110    115    120 

Initial  Gauge  Pressure.Lb.per  Sq.In. 

FIG.  147.     Influence  of  Initial  Pressure  on  the  Economy  of  a  Small,  High-Speed, 
Non-Condensing  Engine. 

above  about  200  pounds  is  at  the  expense  of  increased  first  cost  and 
maintenance  and  is  only  resorted  to  when  small  weight  and  space  are 
the  most  important  considerations. 

The  range  of  pressures  sanctioned  by  modern  practice  for  different 
types  of  engines  is  as  follows: 


Type  of  Engine. 

Range  in  Pres- 
sure (Gauge). 

Average. 

Simple  slow  speed  

60-120 

90 

Simple  high  speed  

70-125 

100 

Compound  high  speed,  non-condensing  
Compound  high  speed   condensing 

100-170 
100-160 

130 
125 

Compound  slow  speed   condensing 

125-200 

150 

Triple  expansion,  condensing 

140-210 

175 

Quadruple  expansion,  condensing 

125-225 

200 

Steam  Pressure :  Trans.  A.S.M.E.,  4-88,  5-269,  6-572;  Peabody,  Thermo- 
dynamics, p.  248;  Ripper,  Steam  Engine,  p.  306;  Engine  Tests,  Barrus,  p.  258. 

169.  Receiver-Reheaters.  —  The  receivers  between  the  cylinders  of 
multi-expansion  engines  are  frequently  equipped  with  heating  coils 
as  illustrated  in  Fig.  332,  the  function  of  which  is  to  superheat  the 
exhaust  steam  before  delivering  it  to  the  cylinder  immediately  follow- 


288  STEAM  POWER  PLANT  ENGINEERING 

ing,  with  a  view  of  reducing  the  losses  occasioned  by  cylinder  conden- 
sation. The  coils  are  supplied  with  live  steam  under  boiler  pressure 
and  may  serve  to  evaporate  a  portion  of  the  moisture  or  to  actually 
superheat  the  steam  supplied  to  the  following  cylinder.  The  question 
of  the  propriety  of  using  reheaters  is  an  open  one,  since  reliable  data 
relative  to  their  use  are  meager  and  discordant.  The  conditions  under 
which  the  few  recorded  tests  were  made  are  too  diverse  to  warrant 
definite  conclusions.  Some  show  an  appreciable  gain  in  economy, 
others  a  decided  loss.  A  reheater  is  of  little  value  in  improving  the 
thermodynamic  action  of  the  engine,  and  is  probably  a  loss  unless  it 
produces  a  superheat  of  at  least  30  degrees  F.,  and  to  be  fully  effective 
should  superheat  above  100  degrees  F.  (L.  S.  Marks,  Trans.  A.S.M.E., 
25-500.)  The  effectiveness  of  the  reheater  will  evidently  be  increased 
by  the  removal  of  the  greater  portion  of  the  moisture  from  the  exhaust 
steam  before  it  enters  the  receiver.  In  the  5000-horse-power  engine 
at  the  Waterside  Station  in  New  York  it  was  shown  that  both  jackets 
and  reheaters,  either  together  or  alone,  were  practically  valueless 
throughout  the  working  range  of  load.  (Power,  July,  1904,  p.  424.) 
Many  similar  cases  may  be  cited  which  show  no  gain  in  economy  with 
the  use  of  the  reheaters.  On  the  other  hand,  with  properly  propor- 
tioned reheaters,  the  gain  may  be  considerable  and  particularly  with 
superheated  steam.  Practically  all  European  engines  operating  with 
highly  superheated  steam  are  equipped  with  receiver-reheaters.  In  all 
cases  the  reheater  effects  a  great  reduction  in  the  condensation  in  the 
low-pressure  cylinders,  but  the  resulting  gain,  considering  the  conden- 
sation in  the  reheater  coils,  may  be  little  if  any. 

In  triple  expansion  pumping  engines  receiver-reheaters  are  found  to 
effect  an  appreciable  gain  in  economy,  and  practically  all  such  engines 
are  equipped  with  them.  In  electric  traction  work  or  where  the  load 
is  a  widely  fluctuating  one  the  reheater  has  been  virtually  abandoned. 
Apart  from  the  consideration  of  fuel  economy,  all  tests  show  a  marked 
increase  in  the  indicated  power  of  the  low-pressure  cylinder  (5  to  15 
per  cent),  and  to  that  extent  it  increases  the  capacity  of  the  entire 
engine.  (G.  H.  Barrus,  Power,  September,  1903,  p.  516.) 

Receivers:  Trans.  A.S.M.E.,  1-174,  178,  9-549;  Ewing,  Steam  Engine,  p.  222J 
Holmes,  Steam  Engine,  p.  456;  Spangler,  Steam  Engineering,  p.  249;  Whitham, 
Steam  Engine,  p.  395;  Power,  June,  1896,  p.  20;  ibid.,  Nov.,  1905,  p.  684. 

Receiver-Reheaters:  Trans.  A.S.M.E.,  1-178,  17-509,  25-443;  Power,  Sept.,  1903, 
p.  516;  Am.  Elecn.,  Oct.,  1902,  p.  480;  ibid.,  July,  1904,  p.  328;  Eng.  Rec.,  May 
9,  1903,  p.  496;  Engineering,  Aug.  8,  1902,  p.  197;  ibid.,  Aug.  21,  1902,  p.  125;  Eng. 
Rec.,  May  28,  1904,  p.  690. 

170.  Jackets.  —  If  the  walls  of  the  cylinder  are  made  double  and 
the  space  between  is  filled  with  live  steam  under  boiler  pressure,  the 


STEAM  ENGINES  289 

cylinder  is  said  to  be  steam  jacketed.  The  function  of  the  jacket  is 
to  reduce  initial  condensation  by  maintaining  the  temperature  of  the 
internal  walls  as  nearly  as  possible  equal  to  that  of  the  entering  steam. 
The  heat  given  up  by  the  jacket  steam,  and  the  resulting  condensa- 
tion, is  usually  a  smaller  loss  than  would  otherwise  result  from  cylinder 
condensation.  However,  tests  of  numerous  engines  with  and  without 
steam  jackets  do  not  agree  as  to  the  conditions  under  which  their 
use  is  profitable,  the  apparent  gain  ranging  from  zero  to  30  per  cent. 
According  to  Peabody,  a  saving  of  5  to  10  per  cent  may  be  made  by 
jacketing  simple  and  compound  condensing 'engines,  and  a  saving  of 
10  to  15  per  cent  by  jacketing  triple  expansion  engines  of  300  horse 
power  and  under.  On  large  engines  of  1000  horse  power  or  more  the 
gain,  if  any,  is  very  small.  (Peabody,  "Thermodynamics,"  p.  400.) 

Other  things  being  equal,  the  smaller  the  cylinder  and  the  lower  the 
piston  speed  the  greater  is  the  value  of  the  jacket.  Experiments 
show  no  advantage  in  increasing  the  jacket  pressure  more  than  a  few 
pounds  above  that  of  the  initial  steam  in  the  cylinder,  and  it  is  usual 
to  reduce  the  pressure  in  the  jackets  of  the  second  and  succeeding 
cylinders  of  multi-expansion  engines.  (Ripper,  "  Steam  Engine,"  p.  170. ) 

To  be  effective,  jackets  should  be  well  drained,  kept  full  of  live  steam, 
and  the  water  of  condensation  returned  directly  to  the  boiler. 

Pumping  engines  and  other  slow-speed  engines  running  at  practi- 
cally constant  load  are  generally  jacketed,  but  in  street-railway  work 
and  in  the  majority  of  manufacturing  plants  carrying  fluctuating  load, 
jackets  are  not  considered  advantageous. 

Whatever  may  be  the  actual  economy  due  to  jacketing,  there  is  no 
question  but  that  the  jacket  greatly  influences  the  action  of  the  steam 
in  the  cylinders,  and  whether  beneficially  or  not  depends  upon  the 
design  and  construction  of  the  engine.  Unless  otherwise  specified, 
manufacturers  usually  build  their  engines  without  jackets. 

Steam  Jackets :  Trans.  A.S.M.E.,  1-175,  190,  2-198,  9-554,  11-141,  149,  328, 
1038,  12-873,  462,  13-176,  14-1356,  15-779,  137;  Golding,  0<f>  Diagram,  p.  39; 
Hutton,  Heat  Engines;  Perry,  Steam  Engine,  p.  369;  Rankine,  Steam  Engine, 
p.  395;  Reeve,  Thermodynamics,  p.  29;  Ripper,  Steam  Engine,  p.  166;  Thurston, 
Manual  of  the  Steam  Engine,  1-598,  622;  Peabody,  Thermodynamics,  p.  322;  Heck, 
Steam  Engine,  p.  123;  Am.  Mach.,  Jan.  30,  1896,  p.  126;  Engr.,  Lond.,  April  21,  1905, 
p.  401 ;  Engineering,  Jan.  30,  1905,  p.  829;  Eng.  Rec.,  April  16,  1898,  p.  423;  Power, 
Feb.,  1898,  p.  17;  ibid.,  Feb.,  1899,  p.  9;  Eng.  Mag.,  June,  1898,  p.  479,  June,  1899, 
p.  496,  Aug.,  1905,  p.  755. 

Increasing  Rotative  Speed:  See  High-Speed  vs.  Low-Speed  Engines,  paragraph  172. 

Compounding  :  See  Compound  Engines,  paragraph  177. 

Reducing  Back  Pressure  :  See  Influence  of  Condensing,  paragraph  179. 

Superheating  :  See  paragraph  181. 


290  STEAM  POWER  PLANT  ENGINEERING 

171.  Single-  and  Double-Acting  Engines.  —  When  steam  pressure  is 
exerted  on  only  one  end  of  the  piston  the  engine  is  said  to  be  single 
acting,  and  when  exerted  alternately  on  one  side  and  the  other  is  said 
to  be  double  acting.     For  high  speed,  minimum  wear  and  tear,  and 
comparatively  cheap  construction  the  single-acting  engine  offers  some 
advantages.     The   Westinghouse   Standard  and  the   Willans  Central- 
Valve  engines  are  typical  of  this  class.     Silent  running  at  high  speed 
is  possible  because  the  pressure  on  the  crank  pin  is  not  reversed.     The 
output  is  only  half  that  of  a  double-acting  engine  of  the  same  size  and 
speed,  but  a  much  higher  rotative  speed  is  permissible,  which  some- 
what offsets  this  disadvantage.    Single-acting  engines  have  been  operated 
successfully  with  speed  as   high  as   1000   r.p.m.,  while  double-acting 
engines  seldom  exceed  350  r.p.m.  and  that  only  for  strokes  less  than 
12  inches. 

Single-Acting  Engines.  —  Comparison  between  Different  Types  of  Engines  :  Trans. 
A.S.M.E.,  2-294.  Economy  of  Single- Acting  Expansion  Engines:  ibid.,  3-252.  Single- 
Acting  Compound  :  ibid.,  12-275.  Steam  Distribution  in  Compound  :  ibid.,  13-557; 
Ewing,  Steam  Engine,  p.  371 ;  Rankine,  Steam  Engine,  p.  478;  Hutton,  Power  Plants, 
p.  77. 

Double-Acting  Engines:  Hutton,  Power  Plants,  p.  73;  Rankine,  Steam  Engine, 
p.  50;  Ewing,  Steam  Engine,  p.  20;  Le  Van,  Steam  Engine,  p.  240. 

172.  High-  and  Low-Speed  Engines.  —  High  rotative  speed  does  not 
necessarily   mean   high   piston  speed.     An   8  x  10   engine   running   at 
300  r.p.m.  has  a  piston  speed  of  only  500  feet  per  minute,  whereas  a 
36  x  72  Corliss  running  at  60  r.p.m.  has  a  piston  speed  of  720  feet  per 
minute.     The  classification  "  high  speed  "  and  "  low  speed  "  refers  to 
rotative  speed  only,  the  former  above  and   the  latter  below  say  150 
r.p.m. 

On  account  of  the  reduction  of  thermodynamic  wastes,  a  high-speed 
engine  should  give  theoretically  a  higher  efficiency  than  the  same  engine 
at  a  lower  speed,  all  other  conditions  being  the  same.  The  effect  of 
speed  upon  economy  is  decidedly  marked  in  engines  and  pumps  taking 
steam  full  stroke.  For  example,  tests  of  a  12  x  1\  x  12  simplex 
direct-acting  steam  pump  at  Armour  Institute  of  Technology  showed 
a  steam  consumption  of  300  pounds  per  I.H.P.  hour  at  10  strokes  per 
minute,  and  only  99  pounds  at  100  strokes  per  minute.  (See  Figs.  274 
and  275.) 

Tests  of  engines  using  steam  expansively,  however,  do  not  furnish 
conclusive  evidence  on  this  point,  some  showing  a  decided  gain  (Pea- 
body,  "  Thermodynamics,"  p.  425),  others  little  or  no  gain  (Barrus, 
"  Engine  Tests,"  p.  260).  For  example,  a  small  Willans  engine  showed 
an  increase  in  economy  of  20  per  cent  in  increasing  the  rotative  speed 


STEAM  ENGINES  291 

from  111  to  408  r. p.m.  (Peabody,  "  Thermodynamics/'  p.  402),  whereas 
the  compound  locomotives  at  the  Louisiana  Purchase  Exposition  showed 
a  loss  in  economy  for  the  higher  speeds  (Publication  by  the  Penn- 
sylvania Railroad  Company).  On  the  other  hand,  a  comparison  of  the 
performances  of  high-  and  low-speed  Corliss  engines  shows  little  differ- 
ence in  economy,  and  a  general  comparison  between  high-  and  low-speed 
engines  furnishes  little  information,  since  nearly  all  high-speed  engines 
are  of  a  different  class  from  the  low-speed  ones.  High-speed  engines  are 
comparatively  small  in  size,  require  larger  clearance  volume,  and  are 
usually  fitted  with  a  single  valve.  Rotative  speed  is  limited  by  design, 
material,  workmanship,  and  cost  of  subsequent  maintenance.  Speeds 
of  400  r.p.m.  and  more  are  not  unusual  with  single-acting  engines, 
whereas  300  r.p.m.  is  about  the  limit  for  double-acting  machines  with 
strokes  over  12  inches  in  length.  A  comparison  of  tests  of  high-speed 
and  low-speed  engines  in  this  country,  irrespective  of  design  and  con- 
struction, shows  the  former  to  be  less  economical  than  the  latter  in 
most  cases.  In  Europe  high-speed  engines  are  developed  to  a  high 
degree  of  efficiency,  and  their  performances  are  comparable  with  the 
best  grade  of  low-speed  engines. 

High-speed  engines  as  a  class  have  the  advantage  of  being  more 
compact  for  a  given  power,  are  simple  in  construction  and  relatively 
low  in  first  cost;  on  the  other  hand,  they  are  subject  to  comparatively 
rapid  depreciation,  excessive  vibration,  and  are  less  economical  in 
steam  consumption. 

High-  and  Low-Speed  Engines.  —  Effect  of  Speed  on  Condensation :  Peabody, 
Thermodynamics,  p.  424.  Effect  of  Speed  on  Economy:  Barrus,  Engine  Tests, 
p.  257;  Trans.  A.S.M.E.,  7-397,  2-198.  Limitation  of  Speed :  Trans.  A.S.M.E., 
14-806.  Effect  of  Speed  on  Economy:  Ripper,  Steam  Engine,  p.  317. 


173-4.  High-Speed  Single-Valve  Simple  Engines.  —  This  style  of 
engine  is  made  in  sizes  varying  from  10  to  500  horse  power.  The 
cylinder  dimensions  vary  from  4x5  to  24  x  24  and  the  rotative  speed 
from  300  to  175  r.p.m. 

When  ground  is  limited  or  costly  and  exhaust  steam  is  necessary 
heating  or  manufacturing  purposes,  the  high-speed  non-condensing 
engine  is  most  suitable  for  horse  powers  of  200  or  less,  being 
compact,  simple  in  construction  and  operation,  and  low  in  first  cost. 
For  sizes  larger  than  this  the  compound  engine  would  probably 
prove  a  better  investment,  except  in  cases  where  fuel  is  very  cheap 
or  large  quantities  of  exhaust  steam  are  to  be  used  for  manufacturing 
purposes. 

Small  high-speed  engines  are  seldom  operated  condensing,  since  the 


•          v 

y  for  \ 

using     1 
beine    / 


292 


STEAM  POWER  PLANT  ENGINEERING 


gain  due  to  reduction  of  back  pressure  is  more  than  offset  by  the  extra 
cost  of  the  condenser  and  appurtenances. 

Engines  are  ordinarily  rated  at  about  75  per  cent  of  their  maximum 
output.  For  example,  a  12  x  12  non-condensing  engine  running  at 
300  r.p.m.,  with  initial  steam  pressure  of  80  pounds  gauge,  is  nor- 
mally rated  at  70  horse  power,  though  it  is  capable  of  developing  90 
horse  power  at  the  same  speed. 

The  steam  consumption  of  high-speed  single-valve  non-condensing 
engines  at  full  load  ranges  from  27  to  50  pounds  per  indicated  horse- 
power hour,  depending  upon  the  size  of  the  unit  and  the  conditions  of 
operation.  An  average  for  good  practice  is  not  far  from  30  pounds. 
With  superheated  steam  a  steam  consumption  as  low  as  18  pounds 
per  horse-power  hour  has  been  recorded. 

Table  39  gives  the  steam  consumption  of  a  number  of  single-valve 
high-speed  engines  running  condensing  and  non-condensing,  and 


50 


100 


1--25 


Per  Cent  of  Bated  Load 


FIG.  148.     Typical  Economy  Curves  of  High-Speed,  Single- Valve,  Non-Condensing 
Engines.     Saturated  Steam. 

Fig.  148  shows  some  of  the  results  for  different  loads.  The  steam 
consumption  is  fairly  constant  from  50  per  cent  of  the  rated  load  to  25 
per  cent  overload,  but  for  earlier  loads  the  economy  drops  off  rapidly. 
The  desirability  of  operating  the  engine  near  its  rated  load  is  at  once 
apparent.  The  curves  show  a  marked  economy  in  favor  of  the  larger 
cylinders,  but  the  engines  are  not  of  the  same  make,  and  the  conditions 
of  operation  are  somewhat  different. 


STEAM  ENGINES 


293 


FIG.  148a.     Assembly  of  Valve  Gear;  Typical  Corliss  Engine. 


Back  Cylinder  Head 
Back  Cylinder  Head  Stud 

Back  Cyl.  Head  Bonnet - 


Corliss  Exhaust  Valve 


,  Steam  pipe 
Steam  Flange 

^Throttle  Valve 
^Planished  Steel  Lagging 
.O)   )  \  Heat  Insulating  Filling 

Corliss  Steam  Valve  Chamber 
Front  Cylinder  Head 
Front  Cylinder  Head  Studs 
"Piston  Rod  Gland  Studs 


liss  Steam  Valve 
ished  Sheet  Steel 
Lagging 

eat  Insulating  Filling 
Steam  Inlet 


Exhaust  Chest' 


^Exhaust  Flange 
""Exhaust  Opening 
Exhaust  Pipe 


Piston  Rod  Gland 


Rod  Packing 


Corliss  Exhaust  Valve 

Sheet  Steel  Lagging 


Heat  Insulating  Filling 


FIG.  148b.     Section  Through  Cylinder;  Typical  Corliss  Engine. 


294 


STEAM   POWER   PLANT   ENGINEERING 


STEAM  ENGINES  295 

The  most  economical  cut-off  for  a  simple  engine  is  about  one-third 
to  one-fourth  stroke  when  running  non-condensing,  and  about  one- 
sixth  when  condensing. 

An  excellent  performance  for  a  small  high-speed  single- valve  simple 
engine  using  saturated  steam  is  shown  by  the  Reeves  piston  valve 
engine  (No.  6,  Table  39)  in  a  test  made  by  Professor  R.  C.  Carpenter. 
When  running  non-condensing  with  initial  steam  pressure  of  114  pounds 
gauge  the  lowest  steam  consumption  was  28  pounds  per  I.H.P.  hour 
and  31.8  pounds  per  B.H.P.  hour.  Referred  to  the  heat-unit  basis, 
this  gives  470  B.T.U.  per  I.H.P.  per  minute,  non-condensing,  and  450 
B.T.U.  per  I.H.P.  per  minute,  condensing. 

A  test  of  a  12  x  12  simple  slide-valve  Buffalo  engine  (No.  10,  Table 
39)  by  Professor  Reeves  gave  30.6  pounds  of  steam  per  I.H.P.  hour, 
initial  gauge  pressure  79.3  pounds.  As  far  as  the  weight  of  steam  and 
B.T.U.  per  I.H.P.  per  minute  are  concerned,  the  Reeves  engine  shows 
the  better  economy,  but  referred  to  the  performance  of  the  ideal  engine 
the  Buffalo  engine  is  the  more  nearly  perfect  of  the  two.  The  latter  con- 
sumed 510  B.T.U.  per  I.H.P.  per  minute,  and  a  perfect  engine  working 
through  the  same  range  in  temperature  would  require  324  B.T.U.  per 
I.H.P.  per  minute;  hence  the  efficiency  ratio  or  the  degree  of  perfection 
of  the  latter  is  324  -*•  510  =  63.5  per  cent  as  against  57.5  per  cent  for 
the  former. 

Locomotive  No.  1499  of  the  Pennsylvania  system  holds  the  record 
for  economy  in  steam  consumption  for  a  single-valve  non-condensing 
engine  (No.  9,  Table  39).  The  steam  consumption  is  23.4  pounds  per 
I.H.P.  hour  (initial  gauge  pressure  196  pounds  per  square  inch),  which 
corresponds  to  a  heat  consumption  of  398  B.T.U.  per  I.H.P.  per  minute. 
So  far  as  the  writer  knows,  this  is  the  best  recorded  performance  for  a 
single-valve  non-condensing  engine  using  saturated  steam. 

The  performance  of  the  Ames  engine  (No.  7,  Table  39)  is  one  of  the 
best  recorded  for  so  low  an  initial  pressure. 

A  small  single-acting  Willans  engine  (No.  1,  Table  39)  holds  an 
exceptional  record  for  economy  for  a  very  small  single-valve  high-speed 
engine,  having  given  a  steam  consumption  of  26  pounds  per  I.H.P. 
hour,  non  condensing,  initial  gauge  pressure  122  pounds,  corresponding 
to  a  heat  consumption  of  436  B.T.U.  per  I.H.P.  per  minute.  All  of 
these  performances  are  at  the  best  rating  of  the  engines.  For  a  con- 
tinually changing  load,  as  in  electric -lighting  service,  the  average  steam 
consumption  is  considerably  greater  than  that  at  full  load  and  depends 
upon  the  "  load  factor"  (the  ratio  of  the  actual  to  the  rated  load). 
This  is  clearly  shown  in  the  curves  of  the  steam  consumption,  Figs.  148 
and  143. 


296 


STEAM  POWER  PLANT  ENGINEERING 


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Barrus,  Engine  Tests,  p.  118 
rrus,  Engine  Tests,  p.  88 
bod  Thermodynamics 


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STEAM  ENGINES 


297 


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visable to  count  on  a  better  steam  consumption  for  this  type  of  engine 
than  30  to  35  pounds  of  steam  per  I.H.P.  hour. 


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FIG.  149. 

Fig.  149  shows  the  effects  of  condensing  on  a  typical  single-valve 
high-speed  engine.  The  gain  in  fuel  economy  may  be  only  an  apparent 
one,  since  the  steam  consumption  of  the  condensing  apparatus  should 
be  rightfully  charged  to  the  engine. 

When  used  in  connection  with  heating  plants  or  manufacturing  plants 
requiring  large  quantities  of  exhaust  steam  the  thermal  efficiency  is 
very  high  and  may  reach  85  per  cent  as  against  22  per  cent  for  the  best 
compound  condensing  engine.  In  general  when  the  requirements  for 
exhaust  steam  are  in  excess  of  the  steam  consumption  of  a  simple  non- 
condensing  engine  a  high-grade  economical  engine  is  without  purpose. 

175.  High-Speed  Multi- Valve  Engines.  —  The  steam  distribution  in 
a  single-valve  engine  may  give  good  economy  for  a  very  small  range  in 
load  but  be  far  from  satisfactory  for  a  wide  range.  This  must  neces- 
sarily be  so  since  admission,  cut-off,  release,  and  compression  are  all 
functions  of  one  valve,  and  any  change  in  one  results  in  a  change  of  the 
others.  To  obviate  the  limitations  of  the  single  valve,  many  builders 
design  engines  with  two  or  more  valves.  With  a  two- valve  engine  cut- 
off is  independent  of  the  other  events,  and  with  four  valves  all  events 
are  independently  adjustable.  In  addition  to  the  flexibility  of  the  valve 
gear,  the  chief  feature  of  the  four-valve  engines  lies  in  the  reduction  of 
clearance  volume  which  is  made  possible  by  placing  the  valves  directly 


298 


STEAM  POWER  PLANT  ENGINEERING 


over  the  ports.  The  valves  may  be  of  the  common  slide-valve  or  rotary 
type.  As  a  class,  four-valve  engines  are  more  economical  than  those 
having  a  less  number  of  valves.  The  advantages  and  disadvantages  of 
the  four-valve  over  the  single-valve  engines  may  be  tabulated  as  below. 
ADVANTAGES.  DISADVANTAGES. 

1.  Better  steam  distribution.  1.   Increased  number  of  parts. 

2.  Better  regulation.  2.    Increased  first  cost. 

3.  Reduced  clearance  volume.  3.    Requires  greater  attention. 

4.  Less  valve  leakage. 

5.  Better  economy. 

The  steam  consumption  of  a  high-speed  four-valve  non-condensing 
engine  varies  from  22  to  35  pounds  of  saturated  steam  per  horse-power 
hour,  with  an  average  not  far  from  27  pounds.  With  superheated  steam 
the  steam  consumption  may  run  as  low  as  18  pounds  per  horse-power  hour. 

An  exceptional  performance  for  a  simple  unjacketed  high-speed  four- 
valve  engine  is  that  of  Engine  No.  17,  Table  39.  With  initial  gauge 
pressure  of  125  pounds  the  steam  consumption  is  22.24  pounds  per 
I.H.P.  hour,  corresponding  to  a  heat  consumption  of  374  B.T.U.  per 
I.H.P.  per  minute. 


Comparative  Economy 

of  a 
(  A  )    Single  Valve  High  Speed 

and  a 

(  B)    Four  Valve  High  Speed 
Non  Condensing  Engine 

15  x  14  Reeves  Simple    (A) 
16  K  16  Flemming  Simple  (  B  ) 


30        30        40 


50        60        70        80        90       100 
Per  Cent  of  Rated  Load 
FIG.  150. 


110      120      130      140 


Fig.  150  gives  a  comparison  between  a  single-valve  and  a  four-valve 
high-speed  engine,  and  though  the  engines  differ  slightly  in  size,  the 


STEAM  ENGINES  299 

conditions  of  operation  were  comparable  and  the  marked  gain  in 
economy  of  the  latter  over  the  former  is  apparent.  Both  perform- 
ances are  exceptional,  and  a  10  to  15  per  cent  greater  steam  consump- 
tion may  be  expected  in  average  good  practice. 

As  a  general  rule  single- valve  simple  engines  do  not  exceed  500  horse 
power  In  size  for  stationary  work,  whereas  1000  horse  power  is  not  an 
uncommon  size  for  the  multi-valve  type. 

High-Speed  Engines,  General  Description:  Trans.  A.S.M.E.,  2-75;  ibid.,  17-117; 
Engr.  U.S.,  Jan.  15,  1903;  Engr.,  Lond.,  April  15,  1904,  p.  379,  April  29,  p.  433, 
May  13,  p.  478,  May  20,  p.  529.  Proportions  of  High-Speed  Engines :  Trans. 
A.S.M.E.,  8-191;  Klein,  High-Speed  Engines;  Button,  Power  Plants,  p.  70,  Thurs- 
ton,  Stationary  Steam  Engines. 

Tests  of  Simple  High-Speed  Engines  :  Am.  Elecn.,  April,  1901,  p.  197,  Dec.,  1903. 
p.  581;  Trans.  A.S.M.E.,  11-723,  18-795;  Elec.  World,  May  20,  1903,  p.  897,  Sept. 
10,  1904,  p.  404,  Oct.  1,  1904,  p.  587,  Feb.  17,  1906,  p.  369;  Engr.  U.S.,  June  1,  1903, 
p.  416,  Nov.  1,  1904,  p.  758;  Engineering,  July  22,  1898,  p.  116;  Eng.  News,  Dec.  3, 
1903,  p.  493;  Eng.  Rec.,  July  6,  1901,  p.  225;  Machinery,  May,  1903,  p.  481;  Power, 
Jan.,  1904,  p.  44,  Nov.,  1904,  p.  651,  Jan.,  1905,  p.  56;  Stevens  Indicator,  Jan., 
1900,  p.  9;  St.  Ry.  Jour.,  Oct.,  1904,  p.  673;  Technology  Quarterly,  Sept.,  1899, 
p.  255. 

176.  Medium  and  Low-Speed  Multi- Valve  Engines.  —  A  comparison 
of  tests  of  high  and  low-speed  single-valve  engines  irrespective  of  design 
and  construction  shows  the  former  as  a  class  to  be  less  economical  than 
the  latter.  With  four-valve  engines  there  is  no  such  disparity,  and  the 
high-speed  type  has  shown  just  as  good  economy  as  the  slow-speed  class. 
For  example,  Engine  No.  17,  Table  39,  with  Corliss  valves  and  a  speed  of 
210  r.p.m.,  gives  practically  the  same  economy  as  Corliss  engine  No.  15 
operating  at  62  r.p.m.  By  far  the  greater  number  of  simple  multi- 
valve  slow-speed  simple  engines  are  of  the  Corliss  type.  They  range 
in  size  from  50  to  3000  horse  power,  with  cylinders  varying  from  12  x  30 
to  48  x  72.  The  smaller  sizes  with  trip-valve  gear  run  at  90  to  100 
r.p.m.,  and  the  larger  at  50  to  75  r.p.m.  Without  the  trip  gear,  speeds 
of  150  r.p.m.  are  not  uncommon,  but  at  this  speed  they  are  usually 
classified  as  high-speed  engines. 

Table  39  gives  the  steam  consumption,  condensing  and  non-con- 
densing, of  a  number  of  four- valve  slow-speed  simple  engines. 

Engine  No.  15  shows  an  unusual  performance  for  a  simple  Corliss 
engine  operating  both  condensing  and  non-condensing.  With  initial 
gauge  pressure  of  103.5  pounds,  the  minimum  steam  consumption  is 
21.5  pounds  per  I.H.P.  hour  for  the  non-condensing  run  and  16.5  pounds 
for  the  condensing  run.  This  corresponds  to  358  B.T.U.  per  I.H.P. 
per  minute,  non-condensing,  and  302  B.T.U.  per  I.H.P.  per  minute, 


300  STEAM  POWER  PLANT  ENGINEERING 

condensing.  The  efficiency  ratios  are  78.0  per  cent  and  53.2  per  cent 
respectively.  The  cylinder  was  jacketed. 

Attention  is  also  called  to  the  record  of  engine  No.  22,  Table  39, 
which  is  of  the  Sulzer  type  with  four  balanced  poppet  valves,  heads 
and  cylinder  barrel  jacketed.  With  79  pounds  initial  pressure  and  a 
vacuum  of  1.36  pounds  absolute,  the  steam  consumption  is  15  pounds 
per  I.H.P.  hour,  corresponding  to  a  heat  consumption  of  275  B.T.U. 
per  I.H.P.  per  minute. 

177.  Compound  Engines.  —  Compound  engines  may  be  divided 
into  three  classes,  tandem,  cross  compound,  and  duplex.  In  the 
tandem  the  two  cylinders  are  end  to  end,  in  the  cross  compound  side 
by  side,  and  in  the  duplex  one  above  the  other.  The  tandem  and 
duplex  compounds  have  the  advantage  of  (1)  compactness  for  a 
given  power,  (2)  less  complication  and  fewer  parts,  and  (3)  low 
first  cost.  The  crank  effort  is  more  variable  than  in  the  cross  com- 
pound. In  very  large  engines  the  low-pressure  stage  is  generally 
divided  between  two  cylinders  of  equivalent  size  to  avoid  an  excess- 
ively large  single  cylinder  and  to  distribute  the  crank  effort.  High- 
speed non-condensing  compounds  are  ordinarily  of  the  tandem  type 
and  are  finding  much  favor  in  isolated  station  work,  as  in  the 
power  plants  of  tall  office  buildings  where  ground  space  is  limited, 
though  the  duplex  compound  is  sometimes  used.  The  .vertical  or 
horizontal  cross  compound  is  generally  installed  in  street-railway 
plants. 

Cylinder  ratios  for  high-speed  single- valve  compound  engines  vary 
from  about  1  to  2£  with  100  pounds  pressure  to  about  1  to  3  with  a 
pressure  of  150  pounds,  and  for  slow-speed  condensing  engines  from  1 
to  3  with  125  pounds  pressure  to  about  1  to  4  with  a  pressure  of  175 
pounds.  G.  I.  Rockwood  recommends  a  ratio  as  high  as  7  to  1,  and  a 
number  of  engines  designed  along  this  line  have  shown  exceptional 
economy.  A  cross  compound  Corliss  engine  at  the  Atlantic  Mills, 
Providence,  R.I.,  with  cylinders  16  and  40x48  (ratio  6.128  to  1) 
gave  the  low  steam  consumption  of  11.2  pounds  of  steam  per  I.H.P. 
hour,  corresponding  to  a  heat  consumption  of  222  B.T.U.  per  I.H.P. 
per  minute.  The  5500-horse-power  engines  of  the  New  York  Edison 
Company  have  a  cylinder  ratio  of  6  to  1.  The  great  majority  of  com- 
pound engines,  however,  have  cylinder  ratios  of  4  to  1  or  less.  The 
8000-horse-power  engines  of  the  Interborough  Rapid  Transit  system 
have  a  ratio  of  4  to  1,  and  the  4000-horse-power  units  of  the  Metro- 
politan Elevated  Company,  New  York,  a  ratio  of  3.5  to  1. 


STEAM  ENGINES 


301 


FIG.  150a.     3500  K.W.  Vertical  Cross-Compound  Corliss  Engine  as  Installed  at  the 
Power  House  of  the  Twin  City  Rapid  Transit  Co.,  Minneapolis,  Minn. 


302 


STEAM  POWER  PLANT  ENGINEERING 


FIG.  150b.     7500  K.W.  Vertical-Horizontal  Double  Compound  Engine  as  Installed  at  the 
59th  Street  Station  of  the  Interborough.     (Manhattan  Type.) 


STEAM  ENGINES 


303 


The  respective  advantages  and  disadvantages  of  compounding  may 
be  tabulated  as  follows: 


ADVANTAGES. 

1.  Permits  high  range  of  expansion. 

2.  Decreased  cylinder  condensation. 

3.  Decreased  clearance  and  leakage 


4.  Equalized  crank  effort. 

5.  Increased    economy     in    steam 

consumption. 


DISADVANTAGES. 

1.  Increased  first  cost  due  to  multi- 

plication of  parts. 

2.  Increased  bulk. 

3.  Increased  complexity. 

4.  Increased  wear  and  tear. 

5.  Increased  radiation  loss. 


The  ratio  of  expansion  for  a  multi-expansion  engine  is  usually 
taken  to  be  the  product  of  the  ratio  of  the  volume  of  large  to  small 
cylinder  divided  by  the  fraction  of  the  stroke  at  cut-off  in  the  high- 
pressure  cylinder.  For  example,  a  compound  engine  with  cylinders 
24,  48  x  48  cutting  off  at  ^  in  the  high-pressure  cylinder  has  a  nominal 
ratio  of  expansion  of  4  -i-  J  =  12.  The  number  of  expansions  at 
rated  load  in  compound  condensing  engines  varies  widely,  ranging 
from  10  to  33,  with  an  average  not  far  from  16. 

The  steam  consumption  shown  by  tests  of  a  number  of  compound 
engines  using  saturated  steam,  condensing  and  non-condensing,  is 
given  in  Table  40.  For  tests  with  superheated  steam  see  Table  43. 


70 


65 
f  60 


20 


Relative  Economy 

of  a 

Simple  and  Compound 

Non-Condensing  High  Speed 

Engine 


20     30 


50      60      70      80      90     100  120 

Developed  Horse  Power  (B.  H.  P.) 

FIG.  151. 


140 


160 


180 


Fig.  151  shows  the  relative  economy  under  comparable  conditions 
of  a  high-speed  simple  and  a  high-speed  compound  engine,  both  run- 
ning non-condensing  and  using  saturated  steam.  The  advantage  of 
the  compound  at  full  load  and  overload  is  very  marked,  though  its 


304 


STEAM  POWER  PLANT  ENGINEERING 


economy  drops  off  rapidly  at  light  loads  and  may  be  less  than  that  of 
the  simple  engine. 

Fig.  152  shows  the  relative  economy  of  two  compound  Corliss  engines 
running  condensing  and  non-condensing,  both  using  saturated  steam. 


A    21,41  x  30  Compound 
B    20,40  x  42  Compound 


£20 
ti 

2  is 


10 


14 


B0 


M 

5    12 


700        800        900        1000       1100       1200       1300      1400 
Indicated  Horse  Power 

FIG.  152. 


It  should  be  borne  in  mind  that  the  object  of  compounding  is  to 
permit  the  advantageous  use  of  high  pressures  and  large  ratios  of 
expansion.  Under  proper  conditions  compounding  may  increase  the 
economy  at  rated  load  about  20  per  cent  for  non-condensing  engines 
and  30  per  cent  for  condensing  engines. 

An  exceptional  performance  of  a  single-valve  high-speed  non-con- 
densing compound  engine  is  that  of  engine  No.  20,  Table  40.  With 
initial  gauge  pressure  of  128  pounds  the  steam  consumption  is  22.3 
pounds  per  I.H.P.  hour,  corresponding  to  a  heat  consumption  of  376 
B.T.U.  per  I.H.P.  per  minute. 

One  of  the  best  performances  of  a  multi- valve  high-speed  compound 
non-condensing  engine  is  that  of  engine  No.  14,  Table  40.  With 
initial  pressure  of  175  pounds  gauge  the  steam  consumption  at  full 
load  is  17.17  pounds  per  I.H.P.  hour,  corresponding  to  a  heat  con- 
sumption of  291  B.T.U.  per  I.H.P.  per  minute. 

The  8000-horse-power  vertical  cross  compound  Corliss  engines  of 
the  Interborough  Rapid  Transit  system  (No.  6,  Table  40),  probably 
hold  the  record  for  economy  for  compound  engines  without  jackets 
and  reheaters,  using  saturated  steam.  With  initial  pressure  of  175 
pounds  gauge  and  absolute  back  pressure  of  2.2  pounds,  the  steam 
consumption  is  11.96  pounds  per  I.H.P.  hour,  corresponding  to  a  heat 


STEAM  ENGINES 


305 


consumption  of  220  B.T.U.  per  I.H.P.  per  minute.  In  estimating  aver- 
age practice  it  would  be  safe  to  add  10  per  cent  or  20  per  cent  to  the 
steam  consumptions  given  in  Table  40. 


Lb. 


1000   1500 


2000   2500   3000 
Gross  K.W.  Output 


80 


3500   4000   4500   5000 


FIG.  153.      Economy  Test  of  the  5500-Horse-Power  Three-Cylinder  Compound  Engine 
and  Generator  at  the  Waterside  Station  of  the  New  York  Edison  Co. 

Fig.  153  illustrates  the  performance  of  the  5500-horse-power  three- 
cylinder  compound  engine  at  the  Waterside  Station  of  the  New  York 
Edison  Company.  The  best  economy  is  11.93  pounds  of  steam  per 
I.H.P.  hour,  corresponding  to  a  heat  consumption  of  221  B.T.U.  per 
I.H.P.  per  minute. 

Compound  Engines. —  Best  Load  for  Compound  Engine  ;  Trans.  A.S.M.E.,  18-674. 
Cylinder  Proportions  for  Compound  and  Triple  Expansion  Engines:  Trans.  A.S.M.E., 
21-1002,  16-762;  Engr.  U.S.,  Sept.  1,  1906,  p.  586;  Eng.  News,  March  2,  1899,  p.  137; 
Eng.  Rec.,  Jan.  7,  1899,  p.  122 ;  Power,  June,  1904,  p.  47.  Laws  of  the  Average 
Simple  vs.  Compound  Engines  under  Variable  Load:  Am.  Mach.,  Sept.  27,  1900,  p.  927; 
Non-condensing  Compound  Engine  for  Office  Buildings:  Eng.  Rec.,  June  18,  1898, 
p.  45.  Economical  Use  of  Steam  in  Non-condensing  Engines:  Eng.  Mag.,  May,  1898, 
p.  213,  July,  1898,  p.  603. 

Compound  Engine  Tests:  Trans.  A.S.M.E.,  24-1274,  25-264;  Engr.  Lond.,  99-546; 
Eng.  News,  Jan.  11,  1906,  p.  44;  Eng.  Rec.,  April  16,  1898,  p.  431,  June  4,  1898, 
p.  1;  Nov.  18,  1899,  p.  579;  Sibley  Jour.,  May,  1901,  p.  346;  St.  Ry.  Jour.,  27-41. 

178.  Triple  and  Quadruple  Engines.  —  Triple  and  quadruple  expan- 
sion engines  are  in  general  use  where  the  load  is  practically  constant, 
as  in  marine  and  pumping-station  practice,  but  have  been  abandoned 
in  street-railway  work  and  in  plants  where  the  load  fluctuates  widely, 


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STEAM  ENGINES  307 

in  favor  of  the  two  or  three-cylinder  compound.  The  best  economy  on 
a  heat-unit  basis  ever  recorded  for  an  engine  using  saturated  steam  was 
that  of  the  Nordburg  quadruple  pumping  engine  at  Wildwood,  Pa., 
which  gave  a  consumption  of  12.26  pounds  per  I.H.P.  hour  and  a  heat 
consumption  of  186  B.T.U.  per  I.H.P.  per  minute  reckoned  above  the 
feed- water  temperature.*  The  Allis  triple  expansion  pumping  engine 
at  Chestnut  Hill  holds  the  record  for  saturated  steam  consumption, 
10  pounds  per  I.H.P.  hour,  and  its  exceptional  performance  of  one 
developed  horse  power  per  1.09  pounds  of  coal  has,  perhaps,  never 
been  excelled.  An  inverted  vertical  marine  cross  compound  engine,  21 
and  36  x  36,  built  by  Cole,  Marchent  &  Morley,  Bradford,  England, 
holds  the  record  for  superheated  steam  consumption,  8.58  pounds  per 
I.H.P.  hour.  (Table  47.)  On  the  heat  basis  (192  B.T.U.  per  I.H.P.  per 
minute),  however,  it  does  not  equal  the  performance  of  the  Nordburg 
engine.  The  above  efficiencies  have  been  exceeded  by  the  binary 
vapor  engine  at  Berlin;  but  this  belongs  in  a  class  by  itself  and  should 
hardly  be  compared  with  the  ordinary  form  of  steam  engine.  (See 
paragraph  182.) 

Triple  Expansion  Engines.  —  Cylinder  Proportions  for  Triple  Expansion  Engines  : 
Trans.  A.S.M.E.,  21-1002,  10-576.  Economy  of  Triple  Expansion  Engines  :  Trans. 
A.S.M.E.,  8-496. 

179.  Effects  of  Condensing.  —  The  effect  of  the  condenser  upon  the 
power  and  economy  of  engines  is  indicated  in  Table  41.  The  curves 
in  Figs.  154  and  155  were  plotted  from  tests  made  by  Professor  R.  L. 
Weighton  on  a  7,  10£,  15J  x  18  triple  expansion  engine  at  Durham 
College  of  Science,  Newcastle-on-Tyne.  The  straight  line  shows  how 
the  mean  effective  pressure  would  vary  with  the  degree  of  vacuum  if 
the  power  increased  directly  with  the  reduction  in  back  pressure. 
The  curved  line  shows  the  actual  M.E.P.,  which  increases  almost  along 
the  theoretical  line  up  to  a  10-inch  vacuum,  from  which  point  on  the 
increase  is  less  marked.  At  26  inches  the  actual  M.E.P.  reaches  an 
apparent  maximum.  These  figures  are  not  applicable  to  all  engines  but 
give  a  good  idea  of  the  limitation  of  the  vacuum  with  the  reciprocating 
engine.  The  gain  in  steam  consumption  due  to  the  condenser  does 
not  indicate  a  corresponding  gain  in  heat  consumption.  For  example, 
Engine  No.  2,  Table  41,  shows  an  apparent  gain  in  steam  consumption, 
due  to  condensing,  of  12.5  per  cent,  the  temperature  of  the  feed  water 
returned  to  the  boiler  being  120  degrees  F.  With  a  suitable  heater 
the  exhaust  of  the  non-condensing  engine  would  be  capable  of  heating 

*  Replaced  in  1905  by  a  Riedler  pumping  engine  on  account  of  high  maintenance 
cost. 


308 


STEAM  POWER  PLANT  ENGINEERING 


X 


X 


Increase  in  Power  Due  to  Vacuum 


Trip. 


e  Expansion  Engine 


10       12       14        16        18       20       22       34 
Vacuum  in  Inches  of  Mercury 

FIG.  154. 


Pounds  of  Steam  per  I.H.P.-Hour 

»•*  H*  »-t  |_l  %_l 

0»  01  -1  00  S 

1 

1 

j  i  1  1  i 

ute.  Above  Hot-well  Temp 

\ 

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Triplb  Expans 

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on  E 

to  Vkcuu 
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310  | 

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Vacuum  in  Inches  of  Mercury 


FIG.  155. 


STEAM  ENGINES 


309 


the  feed  water  to  210  degrees  F.  The  non-condensing  engine  should 
therefore  be  credited  with  210  -  120  or  90  heat  units  per  pound  of 
steam  used,  or,  in  round  numbers,  9  per  cent.  The  difference  between 
12.5  per  cent  and  9  per  cent,  or  3.5  per  cent,  represents  the  net  gain  in 
favor  of  condensing  provided  the  power  necessary  to  create  the  vacuum 
is  ignored.  Actually  the  steam  consumption  of  the  condenser  pumps 
might  be  equal  to  or  greater  than  3.5  per  cent  of  the  steam  generated 
and  the  net  gain  becomes  zero  or  even  negative.  Referring  to  Fig. 
155,  plotted  from  tests  of  the  7,  10J,  15J  x  18  triple  expansion  engine 
mentioned  above,  the  solid  lines  show  the  feed-water  consumption  per 
I.H.P.  hour  and  the  broken  line  the  heat  units  consumed  per  brake 


13.00 


12.50 


12.00 


1LJW 


Rec.in  Lb  .40  Lb  . 
Vac.in  In  -24 


27.5 


FIG.  156.     Performance  of  5500  H.  P.  Engine  at  Waterside  Station  of  New  York 

Edison  Company. 

horse  power  per  minute  measured  above  the  hot- well  temperature. 
The  engine  efficiency,  based  upon  the  water  consumption,  increases  as 
the  vacuum  increases,  reaching  a  maximum  between  26  and  28  inches, 
whereas  the  heat-unit  curve  gives  the  maximum  between  20  and  21 
inches.  Between  22  and  28  inches  the  heat-unit  curve  shows  a  rapid 
falling  off  in  economy.  Tests  of  the  5500-horse-power  engine  at  the 
New  York  Edison  Company's  Waterside  Station  showed  that  increasing 
the  vacuum  from  25.3  to  27.3  inches  decreased  the  water  rate  only 
0.06  pound  per  I.H.P.  (Power,  July,  1904,  p.  424.)  The  results  are 
illustrated  in  Fig.  156.  In  most  cases,  and  particularly  with  large 


310 


STEAM   POWER  PLANT  ENGINEERING 


compound  engines,  the  net  gain  due  to  condensing  is  considerable,  but 
the  feed-water  temperatures  and  power  consumed  by  the  auxiliaries 
should  be  taken  into  account.*  Fig.  149  shows  the  effect  of  vacuum 
on  the  steam  consumption  of  a  small  high-speed  simple  engine,  and 
Fig.  152  of  a  cross  compound  Corliss.  (See  also  paragraph  210.) 

TABLE  41. 

EXAMPLES    OF   THE    EFFECT    OF    CONDENSING    ON    THE    ECONOMY    OF 
RECIPROCATING    ENGINES. 


Non-Condensing. 

Condensing. 

Increase  Due 
to  Condensing. 

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147 

54.7 

19.2 

149 

1.6 

83.4 

14.8 

52.5 

25 

2 

148 

540 

19.3 

147 

4 

16.9 

* 

12.5 

3 

126 

83 

23.8 

130 

7.4 

116 

19.1 

39.8 

19.7 

4 

67.6 

209 

28.9 

67 

4.5 

213 

22 

1.9 

23.5 

5 

103.8 

177.5 

22.1 

103.8 

1.2 

155 

16.5 

* 

25.1 

6 

114 

160 

31 

114 

168 

27 

2 

12.9 

7 

96 

120 

23.9 

96 

"4" 

145 

19.4 

20.8 

18.8 

8 

118 

267 

23.24 

119 

4.2 

276.9 

16 

3.7 

31 

9 

75.9 

310 

25.6 

79 

6.4 

336 

20.5 

8.7 

19.9 

10 

62.5 

451 

30.1 

63.6 

7.8 

444 

23 

* 

23.6 

11 

186.7 

40.4 

18.7 

184.6 

1.6 

29.8 

12.7 

* 

32 

*  Cut-off  changed  for  best  economy. 

1.  7,  10J,  15£  x  18  triple ;  Eng.  News,  Aug.  21,  1902,  p.  127. 

2.  17,  27  x  24  Westinghouse  marine,  non-condensing ;  Power,  August,  1903. 

3.  1,  18  x  10  Buffalo  tandem  compound  ;  Elec.  World,  Sept.  10, 1904,  p.  404. 

4.  18  x  30  four-valve  (slide)  ;  Engine  Tests,  Barrus,  p.  88. 

5.  21,  65  x  43.31  Corliss  ;  Peabody's  Thermodynamics,  p.  382. 

6.  12  x  12  Reeves  simple ;  Elec.  World,  Oct.  1,  1904,  p.  587. 

7.  18  x  48  simple  Corliss  ;  Peabody's  Thermodynamics,  p.  354. 

8.  14,  28  x  24  two-valve  (slide) ;  Engine  Tests,  Barrus,  p.  175. 

9.  17  x  24  two-valve;  Engine  Tests,  Barrus,  p.  70. 

10.  28  x  36  Corliss  ;  Engine  Tests,  Barrus,  p.  97. 

11.  Willans  triple  expansion  central  valve  engine ;  Peabody,  Thermodynamics,  p.  406. 

180.  Throttling  vs.  Automatic  Cut-Off.  —  The  action  of  the  gov- 
ernor in  the  throttling  engine  is  shown  by  the  superposed  indicator 
cards  (Fig.  157)  taken  between  zero  or  friction  load  and  maximum 
load.  The  effect  of  throttling  is  to  reduce  the  pressure  during  admis- 
sion, but  does  not  change  the  point  of  cut-off  or  other  events  of  the 
stroke.  The  steam  may  be  partially  dried  or  even  superheated  by 
throttling,  thus  tending  to  reduce  cylinder  condensation.  Initially 
dry  saturated  steam  at  a  pressure  of  125  pounds  gauge  would  be  super- 
*  See  Power,  Feb.  23,  1909,  p.  381. 


STEAM  ENGINES  311 

heated  about  12  degrees  in  expanding  through  a  throttle  to  90  pounds, 
or  if  it  contained  initially  2  per  cent  moisture  would  be  perfectly 
dried  in  expanding  to  40  pounds.  (See  Table  42.)  Friction  through 
the  valve  also  tends  to  dry  the  steam.  Thus  with  very  light  loads 
the  superheat  may  be  decidedly  appreciable.  The  possible  gain  due 


FIG.  157.     Typical  Indicator  Cards.    High-Speed  Throttling  Engine. 

to  decreased  cylinder  condensation  is  to  some  extent  offset  by  incom 
plete  expansion.  The  best  efficiency  for  a  given  load  is  realized  b} 
a  proper  compromise  between  cut-off  and  initial  pressure.  Experi- 
ments made  by  Professor  Dent  on  (Trans.  A.S.M.E.,  2-150)  on  a 
17  x  30  non-condensing  double-valve  engine  showed  the  most  economical 
results  with  J  cut-off  for  90  pounds  pressure,  J  cut-off  for  60  pounds, 
and  -^50  for  30  pounds.  The  average  throttling  engine  does  not  give 
close  regulation,  the  governor  usually  lacking  sensitiveness.  Tests 
show  the  economy  to  be  better  than  that  of  the  automatic  engine  on 
light  loads,  and  the  crank  effort  more  uniform. 


FIG.  158.     Typical  Indicator  Cards.    High-Speed  Automatic  Engine. 

The  indicator  cards  shown  in  Fig.  158  were  taken  from  a  single- 
valve  high-speed  automatic  engine  operating  between  friction  load  and 
maximum  load.  The  mean  effective  pressure  is  adjusted  to  suit  the 
load  by  the  automatic  variation  in  the  cut-off,  the  initial  pressure 
remaining  the  same.  Since  the  cut-off  is  controlled  by  the  action  of 
the  governor  on  the  single  valve,  all  other  events  of  the  stroke  are 


312 


STEAM  POWER  PLANT  ENGINEERING 


likewise  cnanged.     With  a  four-valve  engine  the  variation  in  cut-off 
does  not  affect  the  other  events. 

The  chief  advantage  of  the  automatic  over  the  throttling  engine  lies 
in  its  sensitive  regulation,  and  while,  in  general,  it  gives  a  lower 
steam  consumption  than  the  throttling  engine,  this  is  probably  in 
most  cases  due  to  superior  construction  and  not  to  the  method  of 
governing. 

TABLE  42. 

SHOWING  THE  INITIAL  PER  CENT  OF  MOISTURE  THAT  WILL  BE  EVAPORATED 
IN  THROTTLING  FROM  A   HIGHER  TO  A  LOWER  PRESSURE. 

Based  on  Marks'  and  Davis'  Steam  Tables. 


Final  Pressures. 

I 

nitial  Pi 

•essure  , 

\bsolute 

80 

85 

90 

95 

100 

105 

110 

115 

120 

80   .         ... 

0  13 

0  24 

0.36 

0  45 

0  55 

0  65 

0  74 

0  83 

75  

0.14 

0.26 

0.37 

0.49 

0  59 

0  70 

0  79 

0  88 

0  97 

70 

0  28 

0  40 

0  52 

0  64 

0  74 

0  84 

0  93 

1  03 

1  12 

65  

0.43 

0.55 

0.66 

0.78 

0.88 

0.99 

1.08 

1  18 

1  26 

60 

0  59 

0  71 

0  83 

0  95 

06 

1  16 

1  25 

1  34 

1  44 

55 

0  77 

0  89 

1  01 

1  13 

23 

1  34 

1  44 

1  53 

1  62 

50    .      .   . 

0  97 

1  09 

1  21 

1  33 

43 

1  54 

1  64 

1  74 

1  82 

45    

1  19 

1  32 

1  44 

1  56 

66 

1  76 

1  86 

1  96 

2  05 

40   

1  44 

1.56 

1  68 

1  80 

.91 

2.02 

2  12 

2  24 

2  30 

35...  
30  

1.72 
2.05 

1.85 
2.18 

1.97 
2.30 

2.10 

2.42 

2.20 
2.53 

2.31 
2.64 

2.41 
2.74 

2.51 

2.84 

2.60 
2  93 

25  

2.44 

2.56 

2.69 

2.82 

2.92 

3.03 

3.13 

3  23 

3  32 

20 

2  90 

3  04 

3  16 

3  29 

3  40 

3  51 

3  61 

3  71 

3  80 

15          .  . 

3  51 

3.65 

3.78 

3.90 

4  01 

4  13 

4  23 

4  33 

4  43 

Final  Pressures. 

Initial  Pressure,  Absolute. 

125 

130 

135 

140 

145 

150 

155 

160 

165 

80 

0.91 
1.05 
1.19 
1.34 
1.52 
1.70 
1.90 
2.13 
2.39 
2.68 
3.01 
3.41 
3.88 
4.51 

0.99 
.13 
.27 
.43 
.60 
.78 
.99 
2.21 
2.47 
2.77 
3.10 
3.49 
3.97 
4.60 

.08 
.21 
.36 
.51 
.68 
.86 
2.08 
2.30 
2.55 
2.85 
3.18 
3.58 
4.06 
4.70 

1.15 
.28 
.43 
.59 
.76 
.94 
2.15 
2.38 
2.63 
2.93 
3.26 
3.66 
4.15 
4.78 

1.22 
1.36 
1.50 
1.66 
1.83 
2.02 
2.22 
2.45 
2.71 
3.01 
3.34 
3.74 
4.22 
4.86 

1.29 
1.43 
1.58 
1.73 
1.90 
2.09 
2.30 
2.52 
2.78 
3.08 
3.41 
3.81 
4.30 
4.94 

1.35 
1.49 
1.64 
1.79 
1.96 
2.15 
2.36 
2.59 
2.84 
3.14 
3.48 
3.88 
4.37 
5.01 

1.41 
1.55 
1.70 
1.85 
2.03 
2.21 
2.42 
2.65 
2.91 
3.21 
3.55 
3.96 
4.45 
5.09 

1.48 
1.62 
1.77 
1.93 
2.10 
2.29 
2.50 
2.73 
2.99 
3.29 
3.63 
4.04 
4.53 
5.17 

75                   

70  

65  
60  

55 

50 

45              .... 

40           

35  

30  

25  

20                      ... 

15              

STEAM  ENGINES 


313 


The  following  performances  of  a  Belliss  250-horse-power  high-speed 
condensing  engine  fitted  with  both  automatic  and  throttling  govern- 
ing devices  give  results  decidedly  in  favor  of  the  throttling  engine. 
(Pro.  Inst.  of  Mech.  Engrs.,  1897,  p.  331.) 


Automatic  Cut-Off. 

Throttling. 

Percentage  of  load     .... 

100 
213 
22.5 

62.5 
132 
22.9 

33 

77.8 
28.5 

25 
53 
34.3 

100 
213 
21 

62.5 
132 
21.7 

33 
77.8 
25.6 

25 
53 
28.4 

Electrical  horse  power.  .  .  . 
Steam  per  I.H.P.  hour.  .  .  . 

Some   of  the   comparative   advantages   and   disadvantages   of  the 
automatic  and  throttling  engines  are  as  follows: 


AUTOMATIC. 


1.  Sensitiveness  of  regulation. 

2.  Increased  ratio  of  expansion. 

3.  Low  terminal  pressures. 


THROTTLING. 


Advantages. 


1.  Low  first  cost. 

2.  Crank  effort  more  uniform. 

3.  Reduced  cylinder  condensation. 

4.  Simplicity  of  regulating  device. 


Disadvantages . 


1.  Increased  cylinder  condensation. 

2.  Greater  variation  in  crank  effort. 

3.  Complicated  valve  gear. 

4.  Low  economy  at  very  early  loads. 


1.  Low  ratio  of  expansion. 

2.  High  terminal  pressure. 

3.  Low  initial  pressure  at  early  loads. 


181.  Influence  of  Superheat.  —  (See  also  paragraph  103.)  Table  43 
gives  test  results  for  several  different  types  of  engines  employing  super- 
heated steam.  These  figures  may  be  compared  with  the  perform- 
ances of  engines  using  saturated  steam  as  given  in  Tables  39  and  40. 
A  decided  gain  in  economy  is  shown  in  favor  of  superheat  for  single- 
cylinder  engines.  With  compound  engines  the  advantage  is  not  so 
apparent,  while  triple  expansion  engines  show  the  least  gain.  Tables 
44  to  46  show  the  effect  of  superheating  on  simple,  compound,  and 
triple  expansion  engines.  (Proc.  A.S.M.E.,  September,  1907.)  As  far 
as  steam  consumption  is  concerned,  most  engines  show  greater  economy 
with  superheated  than  with  saturated  steam,  but  the  gain  in  thermal 
efficiency  is  not  so  marked,  and  when  the  economy  is  measured  in 
dollars  and  cents  per  developed  horse  power,  taking  all  things  into 
consideration  the  gain  is  still  further  reduced  and  in  many  cases  com- 
pletely neutralized. 


314 


STEAM  POWER  PLANT  ENGINEERING 


Fig.  159  gives  the  results  of  a  series  of  tests  made  on  a  number  of 
Belliss  &  Morcom  engines  using  superheated  steam.  (Pro.  Inst.  of 
Mech.  Engrs.,  March,  1905,  p.  302.)  The  engines  were  from  200  to 
1500  kilowatts  capacity  and  were  tested  at  full  load.  It  is  noticeable  that 
the  curves  all  converge  to  a  single  point  and  will  meet  at  about  400 


\ 


\ 


Set 


K.W.Output 

Gen.Coupled 
to  Engine 


208 


220 


308 


362 


500 


700 


Load 

at 

Test 


Full 


I.S.P. 


150 


1456 


Full 


V 


100 


300 


200 
Superheat,  Deg.  F. 

FIG.  159.    Effect  of  Superheat  on  Steam  Consumption. 


400 


degrees  F.  The  results  show  that  if  sufficient  superheat  is  put  into 
the  steam  all  engines  of  whatever  size  are  equally  economical.  Fig.  160 
shows  the  relationship  between  degree  of  superheat  and  the  heat 
consumption  at  various  loads  for  a  300-horse-power  Belliss  &  Morcom 
high-speed  triple  expansion  engine.  (Pro.  Inst.  Mech.  Engrs.,  March, 
1905,  p.  303.)  It  will  be  noted  that  the  variation  in  heat  consump- 
tion at  different  percentages  of  load  becomes  less  marked  as  the  degree 
of  superheat  increases.  With  superheat  of  350  degrees  F.  the  heat 
consumption  from  £  load  to  full  load  is  practically  constant. 


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316 


STEAM   POWER  PLANT  ENGINEERING 


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TABLE  45. 

STEAM  AND  COAL  SAVING  IN  A  COMPOUND  ENGINE,  CONDENSING,  OF  250  I.H.P.  WITH  SUPERHEATED  STEAM  AT 
DIFFERENT  TEMPERATURES. 
Press.  10  atm.  =  142.23  Ibs.;  temp,  of  sat.  steam  354  deg.  F.;  cut-off  6  per  cent;  piston  speed  10  ft.  per  sec.;  automatic  cut-off;  4  poppet  or  piston  valves  per  cyl. 

J 

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Sat.  Steam. 

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Cut-off  Constant.  I.H.P.  Variable. 

1 
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150  to  225  deg.  F. 
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(2)  Direct  .. 
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(2)  Indirect. 

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STEAM   ENGINES 


319 


FIG.  159a      3000  H.P.  Sulzer  Engine  Designed  for  Highly  Superheated  Steam. 


320 


STEAM   POWER  PLANT   ENGINEERING 


STEAM  ENGINES 


321 


Table  47  gives  the  results  of  a  test  made  on  a  21,  36  x  36  inverted 
vertical  marine  engine.     (Engr.,  Lond.,  June  2,  1905,  p.  546.) 

TABLE   47. 

PERFORMANCE  OF  21,  36  X  36  INVERTED  VERTICAL  MARINE  CROSS 
COMPOUND  ENGINE. 

(Engr.,  Lond.,  June  2,  1905,  p.  546.) 


Pressure  on  boiler  side  of  throttle  valve, 
gauge       

117.5 

117.5 

117.5 

117 

114.5 

114.5 

Temperature  of  steam  on  boiler  side  of 
throttle  degrees  F 

743 

738 

749 

751 

732 

726 

Degrees  superheat  of  steam  on  boiler 
side  of  throttle,  F  

395 

390 

401 

403 

384 

378 

Temperature  of  steam  at  admission,  F.  .  . 
Degrees  superheat  of  steam  at  admis- 
sion F                                          

601 
253 

590 
242 

569 
221 

580 
232 

558 
210 

550 

202 

Vacuum,  inches  of  mercury,  absolute.  .  . 
IHP                              

3.4 

481.3 

3 
461.1 

2.22 
347.5 

1.82 
333.5 

1.82 
258 

1.7 
145  52 

M  E  P  referred  to  L  P  cylinder  

26 

29.9 

18.8 

18 

14.4 

7.87 

Revolutions  per  minute    

100.6 

100.7 

100.6 

100.7 

100.7 

100.7 

Pounds  of  steam  per  IHP  hour  

9.09 

9.26 

8.88 

8.68 

8.74 

8.58 

B  T  U  per  IHP  per  minute  

192.2 

201.7 

197.6 

194 

194 

192.1 

See  footnote  *  
B.T.U.  per  I.H.P.  per  minute,  perfect 
engine  

187 
142.4 

189 
142.5 

181 
130.2 

179 
126 

179 
128.5 

175 
128 

Thermodynamic  efficiency  

21.4 

21 

21.4 

21.8 

21.8 

22 

Efficiency  ratio 

72 

72 

66 

65 

66 

67 

Equivalent    evaporation   of   saturated 
steam  reckoned  from  hot  well  
Temperature  of  hot  well 

10.63 
102 

10.81 
101 

10.38 

78 

10.07 
71 

10.12 
64 

10.03 
70 

*  B.T.U.  per  I.H.P.per  minute  based  on  latest  (May,  1908)  values  for  specific  heat  of  super- 
heated steam. 

The  relationship  between  the  weight  of  steam  consumed  per  I.H.P. 
hour  and  the  equivalent  heat  consumption  of  a  250-horse-power  tan- 
dem compound  Van  Den  Kerchove  engine  is  illustrated  in  Fig.  161. 

The  performances  of  engines  using  superheated  steam  should  be 
expressed  in  B.T.U.  per  I.H.P.  per  minute  or  the  equivalent,  as  the 
steam  consumption  alone  gives  no  idea  of  the  true  heat  consumption. 

182.  Binary- Vapor  Engines.  —  A  consideration  of  the  Carnot  or 
Rankine  cycles  shows  that  theoretically  the  efficiency  of  the  steam 
engine  may  be  increased  by  raising  the  temperature  of  the  steam 
supplied  or  by  lowering  the  temperature  of  the  exhaust,  that  is  to  say, 
by  increasing  the  range.  Superheated  steam  development  has  prac- 
tically determined  the  upper  limit,  and  economical  practice  indicates  a 
vacuum  of  about  26  inches,  corresponding  to  126  degrees  F.,  as  the 
average  lower  limit  for  most  efficient  results  from  a  commercial  stand- 
point. 


322 


STEAM  POWER  PLANT  ENGINEERING 


GQ     12 


10 


40  60 

Per  Cent  of  Bated  Load. 


FIG.  160.    Effect  of  Superheat  on  Steam  Consumption. 


i..^^^^^^ 

fKXJ 

24o 
•^40 
235 
230 
£25 
*20 
215 
210 
205 
200 
195 
190 

11( 

10 
9 
8 
7 

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£> 

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Influence  of  Superheat  on  Economy 
(250  H.P.  Tandem  Comp'd  Van  Den) 
Kerchove  Engine 

X 

\ 

\ 

100  200  300 

Degrees  of  Superheat,  Fahrenheit 
FIG.  161. 


STEAM  ENGINES 


323 


In  the  binary-vapor  engine  the  working  range  has  been  considerably 
increased  by  substituting  a  highly  volatile  liquid,  as  sulphur  dioxide, 
for  the  water  which  is  ordinarily  used  as  the  cooling  medium  in  the 
surface  condenser. 

The  SO2  in  condensing  the  exhaust  steam  is  itself  vaporized  and  the 
vapor,  under  a  pressure  of  about  175  pounds  per  square  inch,  used 
expansively  in  a  secondary  reciprocating  engine.  The  exhausted  SO2 
is  discharged  into  a  surface  condenser  in  which  it  is  liquefied  by  cooling 
water  much  the  same  as  in  refrigerating  practice  and  used  over  and 


49.8°F 

Circulating 
Water 
Inlet 

62.7  °_ 

Circulating 
Water 
Outlet 

SOjsYaP«_ 

SO   Vaporizer  and  Steam 

r  —  :  r~"            «  Conden3er  n         -  ~ 

187  ff  Absolute  148    F 

•*                          1 

H| 

j       J     Liquid                          1 
f  7'6     S02  Tank                   . 

1             I  ' 
il 

2*-ABSJA                                                     L 

55°Fji                 r 

_,    56.8I.H.P.            S02       1 

J                     "     Cylinder 

L.P.              42.4I.H.P.   |—  1 

J72°F 
E                             J51.6* 

3                1 

^  1     43.2  I.H.P.       L 

Cylinder                                                j- 

3                1      . 

.P.M. 

! 

a        so, 

Condenser           A 

«u»                _22 

Lbsolutt                                1     
rfF                                Ste< 

H.P.               68.6  I.H.P.  P"^ 

J^       !  173^  Absolute 
LnTln-let                                  U3'5E 

<  —              S02  Exhaust 

FIG.  162.     Diagram  of  Binary- Vapor  Engine. 

over  again.  Referring  to  Fig.  162,  which  illustrates  diagrammatically 
a  binary- vapor  engine  at  the  Royal  Technical  High  School,  Berlin: 
A,  B,  and  C  are  the  three  steam  cylinders  of  an  ordinary  triple  expan- 
sion engine  and  D  the  SO2  cylinder.  All  four  cylinders  drive  a  common 
crank  shaft  E.  F  is  a,  high-pressure  surface  condenser  which  acts  as 
a  vaporizer  for  the  SO2  and  a  condenser  for  the  steam.  G  is  a  surface 
condenser  which  serves  to  condense  the  SO2  vapor.  H  is  a  liquid  SO2 
tank.  The  operation  is  as  follows:  Highly  superheated  steam  enters 
the  high-pressure  steam  cylinder  at  I  and  leaves  the  low-pressure 
cylinder  at  J,  just  as  in  any  steam  engine.  The  exhaust  steam  enters 
chamber  F  and  is  condensed  by  the  liquid  SO2  passing  through  the 
coils.  The  condensed  steam  and  entrained  air  are  removed  from  the 
chamber  by  a  suitable  air  pump.  The  steam  in  condensing  gives  up 
its  latent  heat  to  the  liquid  S02  and  causes  it  to  vaporize.  The  SO2 
vapor  passes  from  the  coils  in  chamber  F  to  the  SO2  engine  D  and 
performs  work.  The  exhausted  SO2  vapor  flows  from  cylinder  D  to 
chamber  G,  and  is  condensed  by  cooling  water  flowing  through  a  series 


324 


STEAM  POWER  PLANT  ENGINEERING 


of  tubes.  The  liquid  SO2  is  collected  in  liquid  tank  H  and  thence  is 
pumped  into  the  coils  in  vaporizer  F.  The  approximate  temperatures 
and  pressures  at  different  points  of  the  cycle  are  indicated  on  the 
diagram. 

A  number  of  experiments  made  by  Professor  E.  Josse  in  the  labora- 
tory of  the  Royal  Technical  High  School  of  Berlin  on  an  experimental 
plant  of  about  200  horse  power  gave  some  remarkable  results.  A  few 
of  the  tests  made  with  highly  superheated  steam  gave  the  following 
average  figures: 

I.H.P.  (steam  end) 146.4 

Steam  consumption  per  I.H.P.  hour . . 12.8 

I.H.P.  (SO2  end) 52.7 

Percentage  of  power  of  SO2  engine 35.9 

Steam  consumption  per  I.H.P.  hour  of  combined  engine... 9.43 

When  operating  under  the  most  satisfactory  conditions  a  perform- 
ance of  8.36  pounds  of  steam  per  I.H.P.  hour  was  recorded,  correspond- 
ing to  a  heat  consumption  of  158.3  B.T.U.  per  minute,  which  is  the 
best  recorded  performance  to  date  (1907)  in  the  history  of  steam- 
engine  economy. 


4000 


3500 


3000 


12500 
5  2000 


1500 


1000 


500' 


75     100  150  200  250  300  350 

Horse-power 

FIG.  163.    Cost  of  Simple  High-Speed  Engines. 


400 


SO2  does  not  attack  the  metal  surface  of  the  engine  unless  combined 
with  water,  in  which  case  sulphurous  acid  is  formed.  There  is,  how- 
ever, no  danger  from  this  cause,  since  the  SO2  being  under  greater 
pressure  effectually  prevents  leakage  of  water  into  the  SO2  system. 


STEAM  ENGINES 


325 


The  SO2  cylinder  requires  no  other  lubrication  than  the  S02  itself, 
which  is  of  a  greasy  nature. 


Cost  to  Dollars 

,x 

x 

ir 

X 

/•A 

X* 

^ 

^m< 

^ 

^ 

^Z- 

^ 

X- 

« 

V* 

x 

^ 

x 

s  ( 

_> 

, 

^'J 

^ 

s 

S/t 

^ 

, 

<4 

V 

J* 

'S 

s 

\ 

s 

/ 

/ 

^ 

s 

,/" 

-X 

'/ 

X 

^ 

^ 

k^' 

200  400  600  800  1000  1300  1400  1600  1800  2000  2200  2400  2600 
"Horse-power 

FIG.  164.   Cost  of  High-Speed  Compound  Engines. 


100  2PO  800  400  500  600  700  800  900  1000  1100  1200 


FIG.  165.   Cost  of  Low-Speed  Engines,  Simple  and  Compound. 


Properties  of  S02  ;  Trans.  A.S.M.E.,  25-181  .  Binary-Vapor  Engines  :  Jour.  Frank. 
Inst.,  June,  1903;  Elec.  World  and  Engr.,  Aug.  10,  1901;  U.S.  Cons.  Reports, 
No.  1139,  Sept.  14,  1901;  Engr.  U.S.,  Aug.  1,  1903;  Sib.  Jour,  of  Eng.,  March,  1902. 


326  STEAM  POWER  PLANT  ENGINEERING 

183.  Cost  of  Engines.  —  In  general  the  cost  of  engines  per  horse 
power  diminishes  as  the  size  increases,  but  is  of  course  governed  by  the 
style  and  workmanship.  Average  figures  may  be  expressed  as  follows 
(Engr.  U.S.,  Nov.  15,  1902,  p.  750) : 

Simple  high-speed  engines Cost  in  dollars  =    300  +    8     X     horse  power 

Setting,  high-speed  engines Cost  in  dollars  =      60  +     0.75  X  horse  power 

Compound  high-speed  engines Cost  in  dollars  =  1000  -f  15    X     horse  power 

Simple  low-speed  engines Cost  in  dollars  =  1000  +  10    X     horse  power 

Compound  low-speed  engines Cost  in  dollars  =  2000  +  13     X     horse  power 

Setting,  low-speed  engines Cost  in  dollars  =    500  -f     1.3    X  horse  power 

These  equations  were  deduced  from  the  curves  in  Figs.  163  to  165, 
which  were  plotted  from  the  actual  costs  of  a  large  number  of  engines. 
Rules  for  testing  steam  engines.  —  See  Appendix  C. 


CHAPTER  X. 

STEAM  TURBINES. 

184.   Classification.  —  The  following  outline  gives  a  classification  of 
few  well-known  steam  turbines: 


Single     f  De  Laval. 

Stage.  \  Sturtevant. 

Nozzle 

Impulse 

Expansion. 

Type. 

Curtis. 

Kerr. 

Steam 

Multi- 
Stage. 

Terry. 
Rateau. 

Turbines. 

Hamilton-Holzworth. 

Blade 

Reaction      (  Multi- 
Type.          \     Stage. 

Parsons. 
Schulz. 
Allis-Chalmers. 

Expansion. 

In  the  impulse  type  the  steam  is  expanded  by  suitable  means  before 
doing  useful  work;  that  is,  its  potential  energy  is  first  converted  into 
kinetic  energy.  In  the  reaction  type  the  conversion  is  not  Complete, 
the  expansion  taking  place  partly  before  doing  work  upon  the  wheel 
and  partly  within  the  blades  of  the  wheel  itself.  Thus  the  steam 
gives  up  a  portion  of  its  energy  by  direct  impulse  in  impinging  against 
the  blades  or  buckets  and  the  balance  by  reaction  in  leaving  them. 
The  impulse  type  may  be  either  single-  or  multi-stage,  depending  upon 
the  number  of  divisions  in  which  expansion  takes  place,  but  the 
reaction  type  is  always  multi-stage.  The  single-stage  impulse  machine 
has  one  row  of  buckets  or  vanes  mounted  on  the  periphery  of  a 
revolving  disk  and  one  set  of  stationary  nozzles.  The  peripheral  velocity 
is  very  high,  ranging  in  practice  from  700  to  1400  feet  per  second.  In 
the  multi-stage  impulse  machines  the  expansion  is  divided  between  a 
number  of  stages,  each  one  exhausting  through  suitably  porportioned 
nozzles  into  the  next  succeeding  stage.  The  steam  velocity  is  thereby 
very  much  reduced,  and  the  peripheral  velocity  may  be  considerably 
lower  for  good  efficiency,  ranging  in  practice  from  200  to  600  feet  per 
second.  In  the  reaction  type  the  steam  flows  through  a  large  number 
of  rows  of  blades  alternately  fixed  and  revolving. 

327 


328  STEAM  POWER  PLANT  ENGINEERING 

184a.  General  Elementary  Theory.  —  A  given  weight  of  steam  at  a 
given  pressure  and  temperature  occupies  a  certain  known  volume  and 
contains  a  known  amount  of  heat  energy.  If  the  steam  is  permitted  to 
expand  to  a  lower  pressure  without  receiving  additional  heat  or  giving 
up  heat  to  surrounding  bodies  it  is  capable  of  doing  a  certain  amount 
of  work  which  will  be  the  same  whether  the  expansion  takes  place  in 
the  cylinder  of  a  reciprocating  piston  engine,  a  rotary  piston  engine 
or  the  nozzles  and  blades  of  a  steam  turbine. 

Let          W  =  weight  of  steam,  Ibs.  per  sec. 

E  =  energy  given  up  by  1  pound  of  steam,  ft.-lbs. 
Pl  =  initial  pressure,  Ibs.  per  sq.  in.  abs. 
Pn  =  final  pressure,  Ibs.  per  sq.  in.  abs. 
H^  =  initial  heat  contents  per  lb.,  B.T.U. 
Hn  =  final  heat  contents  per  lb.,  B.T.U. 

Then  the  heat  available  for  doing  useful  work  is 

W  (H.-Hn)  B.T.U. 

If  the  steam  expands  against  a  resistance,  as,  for  example,  the  piston 
of  a  reciprocating  engine,  the  energy  given  up  in  forcing  the  piston  for- 
ward may  be  expressed 

El  =  778-  W  (H,-  Hn}  ft.-lbs.  (72) 

If  the  steam  expands  within  a  perfect  nozzle  the  energy  will  be  given 
up  in  imparting  velocity  to  the  steam  itself,  thus: 

E2=W^-  ft.-lbs.,  (72a) 

.     £9 

in  which  F1  =  velocity  of  the  jet  in  feet  per  second. 

If  the  velocity  of  the  jet  is  retarded  to  Vn  feet  per  second,  as  by 
placing  a  series  of  vanes  in  its  path,  then  the  energy  given  up  to  the 

vanes  (neglecting  all  losses)  is 

V  2  —  V  2 
E  =  W     l  2         •  (72b) 

If  the  jet  is  brought  to  rest  by  the  vanes  (neglecting  all  losses),  then 
Vn  =  0  and  the  energy  given  up  is 

B>-WZ£.  (72c) 

But  E1=E3.     Hence, 


ln 
from  which 


V,  =  223.9  V%  -  Hn*  (73) 

*  For  most  purposes  it  is  sufficiently  accurate  to  make  223.9  =  224. 


STEAM  TURBINES  329 

The  jet  issuing  from  the  nozzle  is  capable  of  exerting  an  impulse 
equal  to  F  upon  any  object  in  its  path,  thus: 

WV 
F  =  -^-l  Ibs.  (74) 

If  A  =  the  area  of  cross  section  of  the  jet  in  sq.  ft.  and  y  =  weight 
of  steam,  Ibs.  per  cubic  foot,  then  W  =  yAV1}  or 


F  =  --±  Ibs.  (74a) 

I/ 

The  reaction,  R,  of  the  jet  against  the  nozzles  is  equal  in  value  and 
opposite  in  direction  to  the  impulse,  or 

B-P-*I*-iW.  (74b) 

9  9 

The  theoretical  horse  power  developed  by  a  jet  of  steam  flowing  at 
the  rate  of  one  pound  per  second  may  be  expressed 

E  V*-Vn2 

H'R  "  550  =  2<7X550' 
in  which          y^  =  injtjai  velocity  of  the  jet,  ft.  per  sec. 

Vn  =  final  velocity  of  the  jet,  ft.  per  sec. 
Steam  consumption  per  horse  power  hour: 

3600 
TF  =  g-p--  (75a) 

Heat  consumption,  B.T.U.  per  horse  power,  per  min.: 

_  W(H,-qn] 

60 
in  which  qn  =  heat  of  the  liquid  at  pressure  Pn 

Impulse  efficiency  of  the  jet  =  equation  (72b)  -f-  equation  (72c). 

V  2          V   2 

Ei  =  ^V       •  (76) 

*   1 

Thermal  efficiency: 

Et  =  H*  ~  Hn-     See  equation  (68).  (76a) 

^i  -  qn 

Efficiency  ratio  or  "  kinetic  "  efficiency: 

Er  =  W(H^H  y    See  equation  (71).  (76b) 

Equations  (72)  to  (76b)  are  general  and  are  applicable  to  all  turbines 
of  whatever  make. 

The  more  important  types  of  turbines  will  be  discussed  separately 
and  an  application  of  above  equations  will  be  made  in  each  specific  case. 


330 


STEAM  POWER   PLANT  ENGINEERING 


STEAM  TURBINES 


331 


185.  The  De  Laval  Turbine.  —  Fig.  166  shows  a  section  through  a 
De  Laval  steam  turbine  and  gear  case  and  illustrates  the  principles  of 
the  single-stage  "  velocity  "  type.  The  turbine  proper,  to  the  right 
of  the  figure,  consists  of  a  high-carbon  steel  disk  W  fitted  at  the 
periphery  with  a  single  row  of  drop-forged  steel  blades  and  inclosed  in 
a  cast-steel  casing.  The  disk  is  secured  to  a  light  flexible  shaft  and  is  of 
such  a  cross  section  that  the  radial  and  tangential  stresses  through- 
out its  mass  are  of  constant  value.  A  flexible  shaft  is  employed  which 
allows  the  wheel  to  assume  its  proper  center  of  rotation  and  thus  to 
operate  like  a  truly  balanced  rotating  body.*  The  shaft  is  supported 
by  three  bearings,  F,  P,  and  N.  N  is  self-aligning  and  carries  the 
greater  part  of  the  weight  of  the  disk.  P  is  a  flexible  bearing,  entirely 
free  to  oscillate  with  the  shaft,  and  its  only  function  is  to  seal  the  wheel 
casing  against  leakage.  The  power  is  transmitted  through  a  steel 
helical  pinion  K  mounted  on  the  extension  of  the  turbine  shaft  X,  to 
two  large  gears  E,  E  at  a  reduction  in  speed  of  about  10  to  1.  The 
blades,  Fig.  167,  are  made  with  a  bulb  shank  and  fitted  in  slots  milled 


FIG.  167.      De  Laval  Blades. 

in  the  rim  of  the  wheel.  The  flanges,  at  the  outer  end  of  the  blades,  are 
brought  in  contact  with  each  other  and  calked  so  as  to  form  a  continu- 
ous ring.  The  inlet  and  outlet  angles  of  the  blades  are  made  alike  and 
are  32  degrees  for  smaller  sizes  and  36  degrees  for  larger  sizes. 

The  operation  is  as  follows:  Steam  enters  the  steam  chest  D,  Figs. 
166  and  168,  through  the  governor  (shown  in  detail  in  Fig.  169)  and 
is  distributed  to  the  various  adjustable  nozzles,  varying  in  number 
from  1  to  15  according  to  the  size  of  turbine.  In  the  earlier  types 
the  nozzles  were  uniformly  distributed  around  the  circumference,  but 
in  the  later  types  are  arranged  in  groups.  As  illustrated  in  Fig.  168, 

*  The  shaft  diameter  for  a  100-H.P.  turbine  is  but  1  inch  and  for  a  300-H.P. 
approximately  ly£  inches. 


332 


STEAM   POWER   PLANT  ENGINEERING 


the  nozzles  are  placed  at  an  angle  of  20  degrees  with  the  plane  of 
the  disk.  The  steam  is  expanded  adiabatically  in  the  nozzles  to  the 
existing  back  pressure  before  it  impinges  at  high  velocity  against  the 


FIG.  168.     De  Laval  Nozzle. 


blades.  After  giving  up  its  energy  the  steam  passes  into  chamber  G, 
Fig.  166,  and  out  through  the  exhaust  opening.  Fig.  169  gives  the 
details  of  the  governor  and  vacuum  valves.  Two  weights  B  are 


FIG.  169.     De  Laval  Governor. 


pivoted  on  knife  edges  A  with  hardened  pins  C  bearing  on  the  spring  D. 
E  is  the  governor  body,  fitted  in  the  end  of  the  gear  wheel  shaft  K, 
and  has  seats  milled  for  the  knife  edges  A.  The  spring  seat  D  is 


STEAM  TURBINES  333 

held  against  pins  A  by  spiral  concentric  springs,  the  tension  on  which 
is  adjusted  by  a  milled  nut  /.  When  the  speed  exceeds  the  normal, 
centrifugal  force  causes  the  weights  to  fly  outward  and  overcome 
the  resistance  of  the  springs.  This  pushes  pin  G  against  bell  crank  L, 
which  in  turn  closes  the  double-seated  valve,  thus  throttling  the 
supply  of  steam.  To  prevent  racing  in  case  the  load  is  suddenly 
removed  the  vacuum  valve  T  is  added  to  the  governor  mechanism.  Its 
operation  is  as  follows:  The  governor  pin  G  actuates  the  plunger  H 
under  normal  conditions  without  moving  the  plunger  relative  to  the 
bell  crank.  In  case  the  load  is  suddenly  removed,  centrifugal  force 
pushes  pin  G  against  bell  crank  L  until  it  reaches  its  extreme  position 
and  the  valve  is  nearly  closed  and  little  steam  enters  the  turbine.  If 
this  does  not  check  the  speed,  plunger  G  overcomes  the  resistance  of 
spring  My  and  H  moves  relative  to  L,  and  its  adjustable  projection  0 
presses  against  valve  stem  T  and  allows  air  to  rush  into  the  turbine 
chamber  through  passage  P. 

The  power  of  the  turbine  depends  upon  the  number  of  nozzles  in 
action,  and  these  can  be  opened  or  closed  by  a  hand  wheel  on  each. 
Each  nozzle  performs  its  function  as  perfectly  when  operating  alone  as 
when  operating  in  conjunction  with  others. 

De  Laval  turbines  are  made  in  sizes  ranging  from  1J  to  300  horse 
power,  condensing  and  non-condensing,  and  are  designed  to  regulate 
within  an  extreme  variation  of  2  per  cent  from  no  load  to  full  load. 

The  speeds  vary  from  10,600  r.p.m.  for  the  largest  size  to  30,000 
r.p.m.  for  the  smallest,  the  gearing  reducing  these  to  900  and  3000 
r.p.m.,  respectively,  at  the  shaft. 

The  diameter  of  the  wheel  varies  from  4  inches  in  the  smallest  tur- 
bine to  30  inches  in  the  largest,  thus  giving  peripheral  velocities  of 
from  520  to  1310  feet  per  second. 

De  Laval  Turbine  :  Prac.  Engr.,  Jan.  1,  1910;  Trans.  A.S.M.E.,  25-1056;  Elec. 
World,  July  29,  1905,  p.  194,  Oct.  26,  1901,  p.  693;  Eng.  Rec.,  Oct.  19,  1901,  p.  371; 
West.  Elecn.,  June  4,  1904,  p.  463;  Machinery,  Oct.,  1904,  p.  6,  Nov.,  1904,  p.  123; 
Electrician,  March  4,  1904;  Power,  Oct.,  1905,  p.  593. 

See  Table  48  for  results  of  tests  of  turbines  of  this  type. 

186.  Elementary  Theory. —  De  Laval  Turbine. —  The  maximum  theo- 
retical power  developed  by  a  jet  of  steam  flowing  through  a  nozzle  is 
dependent  only  upon  the  weight  of  steam  flowing  per  unit  of  time  and 
the  initial  velocity.  Therefore  the  higher  the  initial  velocity  for  a  given 
rate  of  flow  the  greater  will  be  the  power  developed  and  the  higher  the 
efficiency. 

*  In  Europe  De  Laval  turbines  are  made  as  large  as  3750  H.P.  (See  Power 
and  Engr.,  May  10,  1910,  p.  708.) 


334 


STEAM  POWER  PLANT  ENGINEERING 


The  maximum  weight  of  steam  discharged  through  a  nozzle  of  any 
shape  and  for  a  given  initial  pressure  is  determined  by  the  area  of  the 
narrowest  cross  section  or  throat. 

To  obtain  the  maximum  velocity  at  the*exit  or  mouth,  for  a  given  rate 
of  flow,  the  nozzle  should  be  proportioned  so  that  expansion  to  the 
external  pressure  into  which  the  nozzle  delivers  shall  take  place  within 
the  nozzle  itself.  If  expansion  in  the  nozzle  is  incomplete,  sound  waves 
will  be  produced  and  there  will  be  irregular  action  and  loss  of  energy. 
On  the  other  hand,  if  expansion  in  the  nozzle  is  carried  below  that  of 
the  external  pressure  at  the  mouth,  sound  waves  will  be  produced  with 
subsequent  loss  of  energy  even  greater  than  in  the  former  case. 

Experimental  and  mathematical  investigations  indicate  that  the 
pressure  at  the  narrowest  section  of  an  orifice  or  the  throat  of  a  nozzle 
through  which  steam  is  flowing  falls  to  approximately  0.58  of  the  initial 
absolute  pressure  (with  resultant  velocity  of  about  1400  to  1500  feet  per 
second)  and  any  farther  fall  in  pressure  must  take  place  beyond  the 
narrowest  section.  Thus  for  back  pressures  greater  than  0.58  of  the 
initial  (conveniently  takes  as  £),  maximum  exit  velocity  may  be  ob- 


FIG.  170.     Theoretically  Proportioned  Expanding  Nozzle. 


tained  from  orifices  or  nozzles  of  uniform  cross  section  or  with  sides 
convergent.  For  back  pressure  less  than  0.58  of  the  initial  the  nozzle 
must  first  converge  from  inlet  to  throat  and  then  diverge  from  throat  to 
mouth  in  order  to  obtain  maximum  velocity.  Without  the  divergent 
portion  of  the  nozzle  the  jet  will  begin  to  spread  after  passing  the  throat, 
and  its  energy  will  be  given  up  in  directions  other  than  that  of  the 
original  jet. 

Fig.  170  shows  a  section  through  a  theoretically  proportioned  expand- 
ing  nozzle.  The  cross  section  of  the  tube  at  any  point  n  may  be  cal- 
culated by  means  of  equation 


An  - 


(76o) 


STEAM  TURBINES  335 

in  which 

An  =  area  in  square  feet. 

W  =  maximum  weight  of  steam  discharged,  Ibs.  per  sec. 
Sn  =  specific  volume  of  the  steam  at  pressure  Pn. 

For  saturated  steam  Sn  =  xnun, 

in  which  xn  =  quality  of  steam  at  pressure  Pn  after  adiabatic  expansion 

from  pressure  Pr 
un  =  specific  volume  of  saturated  steam  at  pressure  Pn. 

For  superheated  steam,  see  equation  at  bottom  of  page  131. 

":;  •  • 

Vn  =  velocity  of  the  jet,  feet  per  second. 
Vn  may  be  determined  from  equation  (73): 


Vn  =  223.9      H,  -  Hn. 

By  substituting  Hn  =  heat  contents  corresponding  to  pressure 
Pn  =  0.58  Pl  in  equation  (73)  and  (76c)  the  area  at  the  throat  may  be 
readily  determined.  The  cross-sectional  area  for  other  points  in  the 
tube  may  be  determined  in  a  similar  manner  by  assigning  values  of 
Hn  corresponding  to  the  various  pressures. 

In  case  of  a  perfect  nozzle  Hl  —  Hn  represents  the  heat  given  up 
toward  producing  velocity  by  adiabatic  expansion  from  pressure  Pt  to 
Pn.  In  the  actual  nozzle  the  frictional  resistance  of  the  tube  serves  to 
increase  its  dryness  fraction,  but  in  doing  so  it  decreases  the  amount  of 
energy  the  steam  is  capable  of  giving  up  towards  increasing  its  own 
velocity.  If  y  one-hundred  ths  of  the  heat  Hl  —  Hn  is  utilized  in  over- 
coming frictional  resistance,  then  the  resulting  velocity  will  be 

V  =  223.9  V(l  -y)(Hl-Hn).  (76d) 

The  quality  of  the  steam  after  expanding  to  Pn  against  the  resistance 
will  be  higher  by  an  amount 

In  =  increase  in  quality  =  !  -  -  --  —  >  (76e) 

Tn 

in  which  rn  =  heat  of  vaporization  at  pressure  Pw. 

The  curves  in  Fig.  171,  calculated  by  means  of  equations  (76b)  and 
(73),  show  the  relationship  between  velocity,  quality,  pressure  and 
kinetic  energy  for  all  points  in  a  theoretically  perfect  nozzle  expanding 
one  pound  of  dry  steam  per  second  from  an  initial  absolute  pressure  of 
190  pounds  to  a  condenser  pressure  of  one  pound. 

The  curves  in  Fig.  172  are  based  upon  the  experiments  of  Gutermuth 


336 


STEAM  POWER  PLANT  ENGINEERING 


(Zeit.  d.  Ner.  Ingr.,  Jan.  16,  1904)  and  show  the  effect  of  a  few  shapes 
of  nozzles  and  orifices  on  the  actual  weight  of  steam  discharged  for 
various  rates  of  initial  and  final  pressures,  the  smallest  section  of  the 
tube  remaining  constant. 

The  nozzles  of  most  commercial  types  of  steam  turbines  are  made 
with  straight  sides  as  in  Fig.  168,  so  that  only  the  area  at  the  mouth  need 


6000 


4500 


4000 


3000 


I     2500 

9 

J     2000 

» 

>     1500 


1000 


THEORETICAL    DESIGN    OF   A    DIVERGENT    NOZZLE 


100 

90 

80 

70 


0  40,000      80,000      120,000    160,000    200,000    240,000     280,000 

KINETIC    ENERGY    OF    THE    JET,    IN    FOOT    POUNDS 
FlG.    171. 


be  determined  in  addition  to  that  at  the  throat  in  order  to  lay  out  the 
shape  of  the  tube. 

Equations  (73)  and  (76b)  are  general  and  are  applicable  to  steam 
of  any  quality,  wet,  dry,  or  superheated.  For  steam  initially  dry  and 
saturated  Napier's  rule  offers  a  simple  means  of  determining  the  area 
at  the  throat,  thus: 

W  =  ^   for    Pn  =  or  <  \  Pv.  (76e) 

7U  o 


STEAM  TURBINES 


337 


in  which 


W  =  0.029  A0  VPn  (P,  -  Pn)    for    Pn  >  f  Plt 

W  =  maximum  weight  of  steam  discharged,  Ibs.  per  sec. 

A0  =  area  at  the  throat,  sq.  in. 

Pl  =  absolute  initial  pressure,  Ibs.  per  sq.  in. 

Pn  =  absolute  back  pressure,  Ibs.  per  sq.  in. 


.06 


.05 


.04 


.02 


.01 


.06 

.05 
.04 
.03 
.02 
.01 

^ 

4 

V 

s 

—  x 

^ 

\ 

k 

s3 

X 

\ 

\ 

\ 

\\ 

\ 

r~*i 

r~~ 

\ 

\ 

i 

—  > 

V 

\ 

MI 

L_ 

3 

p— 

_  — 

p2 

p- 
i 

f^~ 

5 

\\ 

I 

•    ' 

i 

2 

j 

p— 

i 

-V 

P« 

— 

Pj=132  Lb.  Per  Sq.  In.  Absolute 
Area  of  Orifice  0.0355  Sq.In. 

Ll 

4 

0     .1     .2     .3     .4     .5     .6     .7     .8     .9    1.0    .1    .2     .3     .4     .5     .6     .7     .8     .9    1.0 
Ratio  — P^  PI  Ratio  -  Pg-r  PV 

FIG.  172.     Flow  of  Steam  through  Nozzles. 

Moyer  ("  The  Steam  Turbine/'  1st  Edition,  p.  40)  states  that  the  ratio 
of  the  area  of  a  correctly  proportioned  nozzle  at  the  throat  A0  to  the  area 
at  any  point  An  is  very  nearly  proportional  to  the  ratio  of  the  pres- 
sure at  point  An  to  the  initial  pressure,  or 

T-JT'  (76f) 

•n-n         r n 

The  entrance  to  the  tube  is  rounded  by  any  convenient  curve. 

The  length  of  the  tube  may  be  roughly  approximated  by  the  following 

formula:  

L  =  \/15  A0,  (76g) 

in  which 

L  =  length  between  the  throat  and  mouth,  in  inches. 

AQ  =  area  at  the  throat,  sq.  in. 

Practice  shows  that  the  cross  section  of  a  nozzle,  whether  circular, 
elliptical,  square  or  rectangular  (the  latter  with  rounded  comers),  has 
very  little  influence  on  the  efficiency  provided  the  inner  surfaces  are 
smooth  and  the  ratio  of  the  area  at  the  throat  to  that  of  the  mouth  is 


338 


STEAM  POWER  PLANT  ENGINEERING 


correctly  proportioned.  The  velocity  efficiency  of  a  properly  propor- 
tioned nozzle  with  straight  sides  is  about  95  to  97  per  cent,  corresponding 
to  an  energy  efficiency  of  92  to  94  per  cent,  so  that  it  is  not  considered 
worth  while  to  attempt  to  follow  the  more  difficult  exact  curves. 

Example :  —  Find  the  smallest  cross  section  of  a  frictionless  conically 
divergent  nozzle  for  expanding  one  pound  of  steam  per  second  from  an 
absolute  initial  pressure  of  190  pounds  to  an  absolute  back  pressure 
of  2  pounds  and  find  six  intermediate  cross  sections  where  the  pressures 
will  be  70,  30,  14.7,  8,  4  and  2  Ibs.  respectively.  Compare  the  velocity 
and  energy  of  the  jet  issuing  from  this  nozzle  with  those  of  an  actual 
nozzle  in  which  10  per  cent  of  the  heat  energy  is  lost  in  friction. 

From  steam  and  entropy  tables  we  find  the  values  of  H,  x,  u,  for 
absolute  pressures  corresponding  to  190,  0.58  X  190  =  110,  70,  30,  etc., 
Ibs.  per  square  inch  as  follows  (theoretical  nozzle): 


H. 

jr. 

u. 

S  =  JTU. 

P,  =  190 

1197.3 

1.00 

2.405 

2.394 

P2  =  110* 

1152.6 

0.960 

4.047 

3.878 

P3=  70 

1117.9 

0.932 

6.199 

5.775 

P4=  30 

1057.2 

0.887 

13.75 

12.27 

P5=  14.7 

1011.3 

0.857 

26.78 

22.95 

Po=   8 

947.8 

0.834 

47.26 

39.29 

P7=   4 

935.6 

0.810 

90.4 

73.2 

P*=   2 

899.3 

0.788 

173.1 

137.0 

*  P2  =  0.58Pi  (  =  pressure  at  throat). 


If  entropy  tables  or  charts  are  not  available,  values  Hl  to  H8  and 
xl  to  xs  may  be  determined  as  outlined  in  equations  (66b)  to  (67g). 

The  different  quantities  for  the  theoretical  nozzle  will  be  calculated 
for  the  exit  pressure  Pn  =  P8  =  2  Ibs.  per  sq.  in  absolute. 


E, 


V8  =  223.9  V%  -  H8 

=  223.9  Vl  197.3  -  899.3 
=  3865  feet  per  second. 
=  778  (H,  -  H8] 
=  778  (1197.3  -  899.3) 
=  232,000  foot-pounds. 

WS 
V 
1  X  137 

3865 
=  .0353  square  foot. 


STEAM  TURBINES 


339 


13.56  VA 


13.56  V.0353 
2.54  inches. 

WV* 


9 

=  3865 
~  32.2 
=  120  pounds. 

THEORETICAL    NOZZLE. 


( 

y 

E 

_4 

d 

F 

Quantity  . 

••••I 

Ft.  per  Sec. 

Ft.-Lbs. 

Sq.  Ft. 

Inches. 

Pounds. 

Formula 

C7Q) 

(72) 

(76c) 

(74) 

V  '  *V 

V,  •  4) 

V,  t  u\sj 

\  *  */ 

110 

1,496 

34,767 

.00259 

0.693 

46.4 

70 

1,995 

61,853 

.00269 

0.702 

61.98 

30 

2,650 

107,485 

.00461 

0.919 

82.3 

Pressures 

14.7 

3,053 

144,742 

.00745 

1.1 

94.8 

8 

3,339 

173,207 

.0119 

1.46 

103.7 

4 

3,624 

203,968 

.0202 

1.92 

112.5 

2 

3,865 

232,000 

.0353 

2.54 

120.0 

In  the  actual  nozzle  these  values  will  be  modified  because  of  the 
frictional  losses.     Thus  for  Pn  =  2  Ibs., 

7  =  223.9 


-  y)  (H,  -  HJ 
=  223.9  \/(l  -  0.1)  (1197.3  -  899.3) 
=  3667  ft.  per  sec. 

E8  =  778  (1  -  0.1)  (1197.3  -  899.3)=  208,800  ft.-lbs. 

,  y(Hi-  HS] 


=  0.788  + 


0.1(1197.3  -899.3) 


1021 


=  0.788  +  0.029 
=  0.817. 

Wxjut 

V8 
_  0.817  X  173.1 

3667 
=  0.0386  sq.  ft., 


STEAM  POWER  PLANT  ENGINEERING 

ds  =  2.66  inches. 
WV        3668 


340 

from  which 


These  various  factors  for  all  given  pressures  have  been  calculated  in 
a  similar  manner  and  are  as  follows: 

ACTUAL    NOZZLE. 


Quantities  j 

V 

Ft.  per  Sec. 

E 
Ft.-Lbs. 

x'. 

A 

Sq.  Ft. 

d 
Inches. 

F 
Ft.-Lbs. 

110 

1,420 

31,317 

.9658 

.00275 

0.711 

44.1 

70 

1,893 

55,632 

.9414 

.00286 

0.723 

58.8 

30 

2,515 

98,257 

.9026 

.00493 

0.951 

78.12 

Pressures  • 

14.7 

2,894 

130,050 

.876 

.0080 

1.2 

98.8 

8 

3,168 

155,858 

.856 

.0127 

1.53 

98.4 

4 

3,438 

183,581 

.836 

.0220 

2.01 

106.8 

2 

3,667 

208,800 

.817 

.0386 

2.66 

114.0 

Many  of  these  values  may  be  determined  directly  from  the  Mollier 
or  total  heat-entropy  diagram  as  described  in  Appendix  H;  in  fact,  the 
Mollier  diagram  has  to  all  intents  and  purposes  supplanted  the  steam 
tables  in  this  connection.  For  superheated  steam  the  diagram  is 
extremely  useful  in  avoiding  laborious  calculations. 


W  = 1250 


FIG.  172a.     Velocity  Diagram.     Ideal  Single-stage  Impulse  Turbine. 

Fig.  172a  gives  a  diagrammatic  arrangement  of  the  blades  in  a 
De  Laval  turbine.  The  nozzle  directs  the  steam  against  the  blades  with 
absolute  velocity  V\  and  at  an  angle  a  with  the  plane  of  the  wheel  XX. 
Since  the  wheel  is  moving  at  a  velocity  of  u  feet  per  second,  the  velocity 


STEAM  TURBINES  341 

vl  of  the  steam  relative  to  the  wheel  is  the  resultant  of  Vl  and  u.  The 
angle  ^  between  vl  and  XX  will  be  the  proper  blade  angle  at  entrance. 
If  the  blade  curve  makes  this  angle  with  the  direction  of  motion  of  the 
wheel  no  shock  will  be  experienced  when  the  steam  enters  the  blades. 
For  convenience  in  construction  the  exit  angle  /?2  is  made  the  same  as 
the  entrance  angle  /?r  Neglecting  frictional  losses  in  the  blade  channels 
the  relative  exit  velocity  will  be  v2  =  vlt  and  the  absolute  velocity  V2  is 
the  resultant  of  v2  and  u.  The  impulse  exerted  by  the  jet  in  striking 

W 
the  vanes  is  —  vlt  and  its  component   in  the  direction   of   motion   is 

W  W 

—  vl  cos  /?t  =  —  (Vl  cos  a  —  u).     As  the  jet  leaves  the  vanes  the  im- 

i/  i/ 

W  W 

pulse  is  -    —  v2  cos  /?2  =  ---  (V2  cos  y  +  u). 
o  u 

The  total  pressure  acting  on  the  vanes,  or  the  actual  driving  impulse,  is 

W  (  ) 

P  =  —  J  F,  cos  a  —  u  —  (  —  V2  cos  y  +  u)  > 

^    (  ) 

W 

=  —(Vl  cos  a  +  V2  cos  y).  (77) 


Equation  (77)  may  also  be  expressed 

W 
P  =  —  -2(VlCos 

-     *  J/ 

The  resultant  axial  force  or  end  thrust  is 


(77a) 


W 

F  =  -  (Vl  sin  a  -  V2  sin  y).  (77b) 

Evidently  if  a  =  y  and   Vl  =  V2  there  will  be  no  end  thrust,  since 
Vl  sin  a  —  V2  sin  y  will  be  zero. 
The  work  done  is 

W 

pu  =  —  u  (Vl  cos  a  +  F2  cos  y),  (77c) 

J/ 

or,  using  equation  (77a)  in  place  of  (77), 

W 
Pu  =  —  •  2  u  (Ft  cos  a  —  w) 

=  —  •  2  (^  cos  a-w2).  (77d) 

c/ 

By  making  the  first  derivative  equal  to  zero 

j-  j  —  2  (uVl  cos  a  -  u2)  I  =  V\  cos  a  -  2  u  =  0, 
or  w  =      F  cos  a. 


342  STEAM  POWER  PLANT  ENGINEERING 

That  is,  for  any  nozzle  angle  a  the  work  done,  Pu,  has  its  greatest 
value  when  u  =  £  Vt  cos  a  or  7  =  90°,  whence 

Pu=  W  ^cos2a.  (77e) 

The  work  for  any  initial  velocity  Vl  becomes  a  maximum  when  a  =  0 
and  u  =  \  V r  This  condition  can  only  occur  for  a  complete  reversal 
of  jet  and  zero  final  velocity.  Substitute  a  =  0  and  u  =  \  V1  in 
equation  (77d). 

WV  2 
Pu  =  — — ->  which  is  necessarily  the  same  as  equation  (72c). 

In  the  actual  turbine  the  various  velocities  will  be  less  than  those  as 
obtained  on  account  of  the  frictional  resistance  in  the  blades,  and  the 
velocity  diagram  should  be  modified  accordingly. 

Example.  Lay  out  the  blades  (theoretical  and  actual)  for  the  nozzle 
in  the  preceding  example,  assuming  that  the  jet  impinges  against  the 
wheel  at  an  angle  of  20  degrees  and  that  the  peripheral  velocity  is 
1250  feet  per  second. 

Theoretical  Case. 

Lay  off  Vl  =  3865  feet  per  second  in  direction  and  amount  as  shown 
in  Fig.  172a  and  combine  it  with  u  =  1250  feet  per  second;  this  gives 
vlf  the  relative  entrance  velocity  as  2725  feet  per  second  and  /?,  the 
entrance  angle  as  29  degrees. 

Lay  off  v2  =  v1  at  an  angle  /?2  =  /?t  and  combine  with  u;  this  gives 
V2,  the  absolute  exit  velocity,  as  1740  feet  per  second. 

The  theoretical  energy  available  for  doing  work  is 

W 

E  =  fg(Vl*-V2*} 

(38652  -  17402)  =  185,000  ft.-lbs. 


64.4 

The  difference  between  232,000  and  185,000  =  47,000  ft.-lbs.  is  evi- 
dently the  kinetic  energy  lost  in  the  exhaust  due  to  the  exit  velocity. 
The  pressure  exerted  by  the  steam  on  the  buckets  is 

W 

P  =  —  (Vl  cos  a  +  V2  cos  7) 

t/ 

=  5^  (3865  X  0.9397  +  1740  X  0.65166) 

fjflmft 

=  148  pounds. 
The  theoretical  impulse  efficiency  is 

V,2  -  722      38652  -  17402  _  .  7Q7 
T"  3865* 


STEAM  TURBINES 


343 


The  theoretical  horse  power  per  pound  of  steam  flowing  per  second  is 

185,000 


H.P. 


550 


=  336. 


Theoretical  steam  consumption  per  H.P.-hr.  is 
3600 


336 


=  10.7  pounds. 


Actual  Case. 

Proceed  as  in  the  theoretical  case,  using  the  actual  absolute  velocity 
Fj  =  3865  Vl  -  y  =  3865  Vl  -  0.10  =  3667  feet  per  second  in  place 
of  the  theoretical  value  Vl  =  3870.  Lay  off  V\  =  3667  at  an  angle  of 
20°  as  before  and  combine  with  u  =  1250,  Fig.  172b. 

U  =  1250 


14=1250 
FIG.  172b.     Velocity  Diagram  as  Modified  by  Friction  Losses. 

The  resultant  VL  =  2530  is  the  velocity  of  the  jet  relative  to  the  wheel, 
and  the  entrance  angle/?  is  found  to  be  29.7  degrees.  The  relative  exit 
velocity  vz  will  be  less  than  vt  because  of  the  blade  friction. 

Assume  the  loss  of  energy  <j>  from  this  cause  to  be  14  per  cent;  then, 
since  the  velocity  varies  as  the  square  root  of  the  energy, 


=  2530  Vl  -  0.14 
=  2346  ft.  per  second. 

The  resulting  absolute  velocity  V2  is  found  from  the  diagram  to  be 
V2  =  1405  ft.  per  second. 
Since  the  loss  of  energy  in  the  nozzle  is 


and  that  in  the  blade 


2<7 
^2  _  (1  _ 


344  STEAM  POWER  PLANT  ENGINEERING 

the  remaining  energy,  deducting   both  losses  in  the  nozzle  and   the 
blades,  is 


(38652  -  0.1  X  38652  -  0.14  X  25302  -  14052) 


64.4 
=  164,200. 

The  losses  due  to  windage,  leakage  past  the  buckets  and  mechanical 
friction  must  be  deducted  from  these  figures  to  give  the  actual  energy 
available  for  doing  useful  work.  Assuming  a  loss  of  15%  due  to  this 
cause,  the  work  delivered  is 

0.85  X  164,200  -  139,570  ft.-lbs. 

The  efficiency  in  the  ideal  case^was  found  to  be  0.797  and  the  available 
energy  185,000  ft.-lbs. 
The  efficiency,  deducting  the  loss  due  to  friction,  etc.,  is 


The  horse  power  delivered  is 


Steam  consumption  per  horse-power  hour  is 

3600 

-^-  =  14.2  pounds. 

The  heat  consumption,  B.T.U.  per  H.P.,  per  minute  is 
16.4  (1197.3  -  94) 

~60~ 

Assuming  the  revolutions  per  minute  to  be  10,000,  the  mean  diameter 
of  the  wheel  to  give  a  peripheral  velocity  of  1250  ft.  per  second  is 

^=2.39^  or  28.6  inches. 

The  determination  of  the  height  and  width  of  vanes,  clearance  between 
nozzles  and  blades,  etc.,  are  beyond  the  scope  of  this  work  and  the  reader 
is  referred  to  the  accompanying  bibliography. 

Blade  Design  for  De  Laval  Turbines:  Moyer,  Steam  Turbine,  Chap.  IV;  Power, 
Mar.  17,  1908,  p.  391. 

Flow  of  Steam  through  Nozzles:  Jour.  A.S.M.E.  Mid.  Nov.,  1909,  April,  1910, 
p.  537;  Engineering,  Feb.  2,  1906;  Engr.,  Lond.,  Dec.  22,  1905;  Eng.  Rec.,  Oct.  26, 
1901;  Power,  May,  1905;  Eng.  News,  Sept.  19,.  1905,  p.  204;  French,  Steam  Tur- 
bines, Chap.  XI;  Pro.  Inst.  Civ.  Engrs.,  Feb.  2,  1906. 

Design  of  Turbine  Disks:  Engr.,  Lond.,  Jan.  8,  1904,  p.  34,  May  13,  1904,  p.  481; 
Jude,  Theory  of  the  Steam  Turbine,  Chap.  XIII;  Thomas,  Steam  Turbines,  Chap.  VI. 


STEAM  TURBINES 


345 


Steam  Turbine  Efficiency:  Power,  Feb.,  1906,  p.  83;  Jude,  Theory  of  the  Steam 
Turbine,  Chap.  VIII. 

Critical  Velocity  of  Shafting:  Jour.  A.S.M.E.,  June,  1910,  p.  1060;  Power,  Sept., 
1903,  p.  484;  Stodola,  Steam  Turbines,  p.  177;  Jude,  Theory  of  the  Steam  Turbine, 
Chap.  XVI;  French,  Steam  Turbines,  Chap.  XV. 

Tests  of  De  Laval  Turbines:  Eng.  Rec.,  Aug.  2,  1902,  p.  100;  Am.  Elecn.,  Aug., 
1905,  p.  445;  Engr.  U.  S.,  Aug.  1,  1905,  p.  526;  Eng.  and  Min.  Jour.,  Nov.  3,  1904, 
p.  706;  Machinery,  Aug.,  1904,  p.  560;  Eng.  News,  June  19,  1905,  p.  62. 

187.  Terry  Turbine.  —  Fig.  173  shows  a  section  through  a  Terry 
turbine,  illustrating  an  application  of  the  impulse  type  with  two  or 
more  velocity  stages.  The  TO  tor,  a  single  wheel  consisting  of  two  steel 


FIG.  173.    Section  through  Terry  Steam  Turbine. 

disks  held  together  by  bolts  over  a  steel  center,  is  fitted  at  its  periphery 
with  pressed  steel  buckets  of  semicircular  cross  section.  The  inner 
surface  of  the  casing  is  fitted  with  a  series  of  gun-metal  reversing  bucket 
arranged  in  groups,  each  group  being  supplied  with  a  separate  nozzle. 
The  steam  issuing  from  nozzle  N,  Fig.  174,  strikes  one  of  the  buckets, 
B,  on  the  wheel  and,  since  the  velocity  of  the  buckets  is  comparatively 
low,  is  reversed  in  direction  and  directed  into  the  first  one  of  the  revers- 
ing chambers.  The  chamber  redirects  the  jet  against  the  wheel,  from 
which  it  is  again  deflected;  this  is  repeated  four  or  more  times  until  the 
available  energy  has  been  absorbed  by  the  rotor.  Terry  turbines  are 
made  in  a  number  of  sizes  varying  from  2  to  800  horse  power,  and  operate 


346 


STEAM  POWER  PLANT  ENGINEERING 


at  speeds  varying  from  210  feet  per  second  in  the  smaller  machine  to 
260  feet  per  second  in  the  larger.  These  low  speed  limits  compared 
with  the  speed  of  single-stage  De  Laval  turbines  are  made  possible  by 
the  application  of  the  velocity  stage  principle  in  the  use  of  the  reversing 
buckets.  The  rotor  of  the  smaller  machine  is  12  inches  in  diameter  and 
runs  at  4000  r.p.m.,  and  that  of  the  larger,  48  inches,  running  at  1250 


FIG.  174.    Arrangement  of  Buckets  and  Reversing  Chambers  in  a  Terry  Steam  Turbine. 

r.p.m.  Since  the  flow  of  steam  into  and  from  the  buckets  is  in  the  plane 
of  the  wheel  there  is  no  end  thrust. 

For  a  description  of  the  Bliss,  Dake,  Sturtevant  and  Wilkinson  steam 
turbines  with  results  of  tests  see  "  Small  Steam  Turbines,"  by  G.  A. 
Orrok,  Jour.  A.S.M.E.,  May,  1909,  and  contributed  discussion,  Sept., 
1909.  See  also,  "The  Development  of  the  Small  Steam  Turbine," 
Eng.  Mag.,  Dec.,  1908,  and  Jan.,  1909. 

188.  Kerr  Turbine.  —  Fig.  175  shows  a  longitudinal  section  through, 
and  Fig.  176  a  sectional  elevation  of,  a  Kerr  steam  turbine.  This 
turbine  is  of  the  impulse  type  and  built  on  the  principle  of  the 
Pelton  water  wheel,  which  it  resembles  in  many  respects.  The  rotor 
consists  of  a  series  of  steel  disks  R,  R,  Fig.  175,  mounted  on  a  steel 
shaft.  A  series  of  drop-forged  mild  steel  buckets  of  the  double-cup 
type  are  secured  to  the  periphery  and  riveted  in  dovetail  slots.  The 
stator  is  made  up  of  a  number  of  cast-iron  diaphragms  S,  S,  with  cir- 
cular rims,  which  are  tongued  and  grooved  and  when  drawn  together 
form  a  continuous  cylinder.  Square  cold-rolled  steel  nozzle  bodies 
N,  N  are  expanded  and  beaded  in  the  diaphragm  near  the  rim  and 
the  nozzles  screwed  into  them.  The  bearings  B,  B  are  of  the  oil-ring 
type;  no  thrust  blocks  are  necessary,  as  each  element  is  practically 
balanced.  The  operation  is  as  follows:  Steam  enters  the  turbine  at 
inlet  A  and  passes  through  balanced  throttling  valve  V  (controlled  by 
governor  G)  to  the  circular  cored  space  H,  H  extending  around  the 


STEAM  TURBINES 


347 


348 


STEAM  POWER  PLANT  ENGINEERING 


entire  casing.  Space  H,  H  acts  as  an  equalizer  and  insures  uniform 
admission  to  the  first  row  of  nozzles,  where  steam  is  partially  expanded 
and  the  kinetic  energy  imparted  to  the  rotor  through  the  medium  of 
the  buckets.  The  steam  leaves  the  buckets  at  practically  zero  velocity 


FIG.  176.     Kerr  Steam  Turbine;  Sectional  End  Elevation. 

and  is  again  expanded  through  the  second  set  of  nozzles.  This  process 
is  repeated  in  each  stage  and  the  exhaust  steam  leaves  the  turbine 
at  0.  Fig.  177  shows  a  diagrammatic  arrangement  of  the  governor. 
The  governor  weight  is  turned  from  solid  steel  and  split  into  two  piece? 


<  Section  A-A 

FIG.  177.     Kerr  Steam  Turbine  Governor. 


of  semi-cylindrical  form  with  the  center  of  gravity  near  the  center  of  the 
shaft.  The  weights  are  supported  at  three  points.  The  hardened  steel 
knife  edge  at  B  is  of  sufficient  length  for  the  stresses  involved.  The 
curve  of  rolling  contact  C  is  such  that  the  bearing  between  the  weigh' 


STEAM  TURBINES 


349 


FIG.  178.     Four-stage  Vertical  Curtis  Turbo-Generator.     Base  Condenser  Type. 


350  STEAM   POWER  PLANT  ENGINEERING 

and  the  cam  collar  is  always  on  the  line  of  centers.  The  outward  move- 
ment of  the  weights  compresses  the  spring  and  operates,  through  lever 
connections,  the  balanced  piston  valve  controlling  the  flow  of  steam. 
The  movement  of  the  center  of  gravity  is  indicated. 

The  Kerr  turbine  is  very  simple  in  design,  compact,  noiseless,  and 
low  in  cost  of  repairs.  Its  performance  compares  favorably  with  all 
other  types  of  turbines  of  similar  size  and  capacity. 

An  18-inch  Kerr  turbine  direct  connected  to  a  multi-stage  Worthing- 
ton  centrifugal  pump  at  the  Armour  Institute  of  Technology  gives  a 
steam  consumption  when  running  non-condensing  comparable  with 
that  of  high-grade  non-condensing  engines. 

Rateau  Turbine:  Trans.  A.S.M.E.,  25-782;  Eng.  Mag.,  Oct.,  1903,  p.  49;  St.  Ry. 
Jour.,  April  18,  1903. 

Zolly  Steam  Turbine  :  Elec.  Rev.,  Sept.  2,  1904. 

189.  The  Curtis  Steam  Turbine.  —  Figs.  178  to  183  show  the  general 
arrangement  and  a  few  details  of  the  Curtis  steam  turbine,  which  is  of 
the  compound  or  multi-stage  velocity  type.  The  total  expansion  is 
carried  out  in  one  or  more  compartments  or  stages,  each  stage  compris- 
ing a  set  of  expanding  nozzles  and  a  wheel  carrying  two  or  more  rows  of 
buckets.  A  high  initial  velocity  is  given  to  the  jet  in  each  stage  by 
expansion  in  the  nozzles  as  in  the  De  Laval,  and  the  energy  absorbed  by 
successive  action  upon  the  series  of  moving  and  stationary  vanes 
arranged  somewhat  as  in  the  Parsons  turbines,  paragraph  192.  In  the 
latter,  however,  the  difference  in  pressure  between  the  two  sides  of  each 
vane  induces  flow  by  continuous  expansion,  while  in  the  former  the 
moving  vanes  in  any  one  stage  simply  absorb  the  kinetic  energy  already 
created  by  expansion  in  the  nozzle.  The  action  is  as  follows :  Steam  enters 
stage  (1),  Fig.  180,  through  the  first  set  of  nozzles,  and  is  partially 
expanded.  With  the  resulting  initial  velocity  it  impinges  against  the 
first  row  of  moving  blades  and  gives  up  part  of  its  energy,  and  is  then 
deflected  through  the  adjoining  stationary  blades  to  the  next  set  of 
moving  vanes,  where  its  velocity  is  still  further  reduced,  and  so  on  until 
it  has  been  brought  practically  to  rest.  From  this  stage  the  steam 
flows  at  reduced  pressure  through  nozzles  of  stage  (2),  which  are  suffi- 
cient in  number  and  in  size  to  afford  the  greater  area  required  by  the 
increased  volume.  In  expanding  in  these  nozzles  it  acquires  new 
velocity  and  gives  up  energy  to  the  moving  blades  as  before.  This 
process  is  repeated  through  two  to  five  stages,  depending  upon  the  size 
of  turbine.  Fig.  178  shows  a  partial  section  of  a  four-stage  5000-kilo- 
watt  machine.  R,  R  are  sections  through  the  revolving  wheels,  which 
in  this  particular  turbine  are  nine  feet  in  diameter  and  keyed  to  the 


STEAM  TURBINES 


351 


852 


STEAM  POWER  PLANT  ENGINEERING 


vertical  shaft  S.  On  the  periphery  of  each  wheel  are  bolted  two  rows 
of  blades  or  vanes,  with  a  stationary  or  intermediate  row  attached  to 
the  casing  between  them.  The  buckets  are  made  of  rolled  nickel  bronze, 
hammered  to  shape  and  finish.  The  .roots  are  dovetailed  into  the  holders 
and  the  tips  are  tenoned  and  riveted  into  a  shroud  ring,  thus  insuring 
positive  spacing  and  a  rigid  construction.  Between  each  pair  of  wheels 
is  a  stationary  steam-tight  diaphragm  P,  which  contains  the  nozzles 
through  which  the  steam  is  expanded  from  the  preceding  stage.  It  will 
be  noticed  that  the  buckets  and  nozzles  increase  rapidly  in  size  in  suc- 
ceeding stages  as  the  pressure  falls  and.  the  volume  of  steam  increases. 


\  D  D  I  D  D  I    TT 


STAGE 

No.  2   5 


MOVING  BLADES' 

STATIONARY  BLADES 

MOVING  BLADES 

NOZZLE  DIAPHRAGM 

MOVING  BLADES 

STATIONARY  BLADES 
MOVING  BLADES 


EXHAUST 

FIG.  180. 


The  parts  are  so  proportioned  that  the  steam  gives  up  approximately 
-of  its  energy  in  each  stage,  n  representing  the  number  of  stages. 

The  number  of  stages  and  the  number  of  vanes  in  a  stage  are  governed 
by  the  degree  of  expansion,  the  peripheral  velocity  which  is  desirable 
or  practicable,  and  by  various  conditions  of  mechanical  expediency. 
The  number  of  admission  valves  vary  in  number  and  in  location  with 
the  size  of  turbine.  The  automatic  stage  valve  G  connects  the  first 
stage  directly  to  a  set  of  auxiliary  second-stage  nozzles.  Thus  the 
overload  capacity  is  increased  by  widening  the  steam  belt  and  not  by 
admitting  high-pressure  steam  into  an  intermediate  stage  as  was  for- 
merly the  practice  with  Curtis  turbines.  This  method  of  overload  con- 
trol results  in  higher  efficiency  than  with  the  older  system. 


STEAM  TURBINES 


353 


Curtis  turbines  appear  to  have  a  wider  range  of  economical  application 
than  any  other  type,  commercial  sizes  ranging  from  a  small  horizontal 
unit  of  7  kilowatts  rated  output  to  vertical  units  of  20,000  kilowatts 
capacity  on  the  continuous  24-hour  basis.  The  smaller  machines, 
1000  kilowatts  and  under,  are  usually  of  the  horizontal  type,  and  the 
larger  units,  3500  kilowatts  and  larger,  are  of  the  vertical  type.  Be- 
tween 500  and  3500  kilowatts  they  are  made  both  vertical  and  horizontal. 
All  Curtis  turbines  are  governed  by  "  cutting-out  nozzles";  that  is, 
full  initial  pressure  is  maintained  in  all  the  nozzles  that  are  open  and 


FIG.  181.     Section  through  Curtis  Governor. 

the  capacity  of  the  machine  is  controlled  by  varying  the  number  in 
operation.  Units  under  1500  kilowatts  are  ordinarily  controlled  by  a 
mechanical  valve  gear  and  the  larger  units  by  an  indirect  or  relay  system. 
In  the  older  types  this  relay  system  was  electrically  operated;  in  the 
modern  machines  the  valves  are  hydraulieally  controlled. 

Fig.  181  shows  a  section  through  a  typical  Curtis  governor.  Speed 
regulation  is  accomplished  by  the  balance  maintained  between  the 
centrifugal  force  of  moving  weights  A  A  and  the  static  force  exerted  by 
spring  D.  The  governor  is  provided  with  an  auxiliary  spring  F,  for 
varying  its  speed  when  synchronizing,  the  tension  in  which  is  varied  by 


354 


STEAM  POWER  PLANT  ENGINEERING 


a  small  pilot  motor  controlled  from  the  switchboard.     The  movement 
of  the  governor  weights  is  transmitted  through  rod  C  to  arm  H  and  by 
means  of  the  latter  to  the  controlling  mechanism  of  the  valve  gear. 
Fig.  182  gives  an  assembly  view  of  the  mechanical  valve  gear  as 


FIG.  182.     Assembly  of  Mechanical  Valve  Gears  for  300-Kw.  Curtis  Steam  Turbine. 

applied  to  a  300-kilowatt  unit.  The  valve  stems  extend  upward  through 
ordinary  stuffing  boxes  and  are  attached  to  notched  crossheads  8,  8. 
Each  crosshead  is  actuated  by  a  pair  of  reciprocating  pawls  or  dogs, 
6,  6,  the  lower  one  of  which  closes  the  valve  and  the  upper  one  opens  it. 
The  several  pairs  of  pawls  are  hung  on  a  common  shaft  which  receives 
a  rocking  motion  from  a  crank  driven  by  the  turbine  shaft.  The  cross- 


STEAM  TURBINES 


355 


heads  have  notches  milled  in  the  side  in  which  the  pawls  engage  to  open 
or  close  the  valve,  the  engagement  being  determined  by  shield  plates  2, 
the  positions  of  which  are  controlled  by  the  governor  through  the  medium 
of  suitable  levers.  Shield  plates  6  are  set  one  a  little  ahead  of  the 
other  to  obtain  successive  opening  or  closing  of  the  valves.  The 
pawls  are  held  in  position  when  not  in  contact  with  the  shield  plates  by 
springs  W. 

Fig.  183  gives  a  diagrammatic  arrangement  of  the  hydraulically 
controlled  valve  gear  mechanism.  The  motion  of  governor  g  is  trans- 
mitted through  lever  i  to  lever  a  of  the  pilot  valve  /.  Pilot  valve  / 
controls  the  supply  of  oil  (under  pressure)  in  cylinder  k  the  piston  of 
which  actuates  rods  Z,  I  The  movement  of  rod  I  is  transmitted  through 


Turbine 


FIG.  183.     Diagrammatic  Arrangement  of  Hydraulically  Operated  Valve  Gear, 

Curtis  Turbine. 

rack  m  to  a  small  pinion.  This  pinion  is  mounted  on  the  end  of  a  shaft 
fitted  with  a  number  of  cams,  one  a  little  ahead  of  the  other,  each  cam 
controlling  the  opening  and  closing  of  a  steam  valve  through  the  medium 
of  rocker  arm/.  As  the  load  on  the  turbine  increases  the  governor  slows 
down  and  causes  the  cam  shaft  to  rotate  in  a  reverse  direction  indicated 
by  the  arrow  points  in  Fig.  183.  This  causes  a  proportionate  number 
of  valves  to  be  lifted  and  held  open,  the  number  increasing  as  the  load 
increases,  until  all  are  open.  Should  the  load  continue  to  increase,  as 
in  the  case  of  overload,  the  secondary  valve  opens  as  previously  described, 
connecting  the  first  stage  with  a  set  of  auxiliary  second  stage  nozzles. 
Only  the  nozzles  in  the  first  stage  are  controlled  by  the  governor. 
Should  the  turbine  run  above  normal  speed  the  emergency  stop  valve 
automatically  closes  the  admission  of  steam  to  the  nozzles.  This  device 


356 


STEAM  POWER  PLANT  ENGINEERING 


consists  of  a  steel  ring  placed  around  the  shaft  between  the  turbine  and 
the  generator.  This  ring  is  eccentrically  mounted  and  the  unbalanced 
centrifugal  force  is  balanced  by  a  helical  spring.  When  the  predeter- 
mined speed  is  reached  the  centrifugal  force  overcomes  the  spring  ten- 
sion and  the  ring  moves  in  a  still  more  eccentric  position.  In  this  posi- 
tion the  ring  strikes  a  bell  crank  lever  which  trips  the  throttle  valve 


0  12345678 

Steam  Belt  Area 

FIG.  183a.     Steam  Belt  Area  in  Five-Stage  Curtis  Turbine. 


and  permits  it  to  close  by  its  own  weight  and  the  unbalanced  pressure 
on  the  valve  stem. 

In  the  ;Curtis  turbine  the  area  of  the  steam  admission  is  limited  to  a 
small  portion  of  the  circumference  in  the  first  stages  and  does  not 


OIL    DRAIN 
OIL  SUPPLY 


FIG.  183b.     Step  Bearing  for  Curtis  Turbine. 

extend  around  the  entire  circumference  until  the  last  stage  is  reached. 
See  Fig.  183a. 

The  step  bearing  of  a  vertical  machine  is  illustrated  in  Fig.  183b.  The 
weight  of  the  rotor  is  supported  by  oil  under  pressure  forced  between 
the  bearing  blocks  M  and  P,  thus  permitting  the  shaft  S  to  revolve 
on  a  film  of  oil.  The  smaller  disk  M  is  attached  by  dowels  F  to  the 
main  shaft.  Carbon  packing  rings  0,  0  are  used  above  the  bearing  to 
prevent  leakage,  and  adjustment  is  provided  in  the  lower  bearing  block 


STEAM  TURBINES 


357 


by  means  of  set  screws.  The  oil  pressure  varies  from  150  to  750  pounds 
per  square  inch  according  to  the  size  of  machine,  the  higher  pressures 
being  used  in  the  larger  machines. 

Fig.  183c  gives  a  diagrammatic  outline  of  the  oiling  system.  A  tank,  of 
sufficient  capacity  to  contain  all  the  oil  and  fitted  with  suitable  straining 
devices  and  a  cooling  coil,  is  located  at  a  level  low  enough  to  receive 
oil  by  gravity  from  all  points  lubricated.  A  pump  draws  oil  from  this 
tank  and  delivers  it  at  a  pressure  about  25  per  cent  higher  than  that 
required  to  sustain  the  weight  of  the  turbine  in  the  step  bearing.  A 
spiral  duct  baffle  connects  the  source  of  pressure  to  the  step  bearing  and 
serves  to  regulate  the  oil  supply  to  the 
lower  end  of  the  shaft.  This  source  of 
pressure  is  also  connected  through  a 
reducing  valve  to  the  upper  oiling  sys- 
tem of  the  machine,  in  which  a  pressure 
of  about  60  Ibs.  to  the  square  inch  is 
maintained.  This  system,  which  includes 
a  storage  tank  partly  filled  with  com- 
pressed air,  operates  the  hydraulic  gov- 
ernor mechanism  and  supplies  oil  to  the 
upper  bearings.  Delivery  of  oil  to  these 
bearings  is  regulated  by  adjustable 
baffles  designed  to  offer  resistance  to 
the  oil  flow  without  forcing  the  oil  to 
pass  through  any  very  small  opening 
which  might  easily  become  clogged.  A 
relief  valve  is  provided  to  prevent  the 
pressure  in  the  upper  part  of  the  oiling 
system  from  rising  above  a  desirable 
limit.  Drain  pipes  from  the  upper  bear- 
ings and  from  the  hydraulic  cylinder  and  relief  valve  all  discharge  into 
a  common  chamber,  in  which  the  streams  are  visible,  so  that  the  oil 
distribution  can  always  be  easily  observed.  At  some  point  in  the  high- 
pressure  system  adjacent  to  the  pump  it  is  desirable  to  install  a  device 
to  equalize  the  delivery  of  oil  from  the  pump,  as  is  done  by  the  air  cham- 
ber commonly  used  with  pumps  designed  for  low  pressure.  A  small 
spring  accumulator  is  furnished  for  this  purpose,  except  in  cases  where 
weighted  storage  accumulators  are  used.  In  large  stations  where 
several  machines  are  installed,  a  storage  accumulator  is  desirable  and 
can  advantageously  be  so  arranged  that  it  will  normally  remain  full, 
but  will  discharge  if  pressure  fails,  and  in  doing  so  will  start  auxiliary 
pumping  apparatus. 


TO     ACCUMULATOR  ^ 


FIG.  183c.     Arrangement  of  Oiling 
System  for  Curtis  Turbine. 


358 


STEAM  POWER  PLANT  ENGINEERING 


DIRECT    CURRENT. 


Kw. 

R.p.m. 

Kw. 

R.p.m. 

15 

4,000 

150 

2,000 

25 

3,600 

300 

1,800 

75 

2,400 

500 

1,500 

ALTERNATING    CURRENT. 


300 

1,800 

2,000 

900 

500 

1,800 

3,000 

\ 

1,000 

1,200 

to 

600-750 

1,500 

900 

20,000 

] 

For  the  description  of  a  typical  steam  turbine  station  equipped  with 
Curtis  turbines  see  Chapter  XX. 

General  Description  of  Curtis  Turbines:  Power,  March,  1909;  Engr.  U.  S.,  Jan.  1, 
1908,  p.  115;  Power  &  Engr.,  Feb.  25,  1908,  p.  284,  Feb.  25,  1908,  March  3,  1908; 
Elec.  Wld.,  June  17,  1905,  p.  1136. 

Guide  Bearings,  Oil  Distribution  and  Carbon  Packing:  Power  &  Engr.,  April  14, 
1908. 

Mechanical  Valve  Gear:  Power  &  Engr.,  March  10,  1908,  p.  356. 

Hydraulic  Valve  Gear:  Power,  March,  1909,  p.  189. 

190.  Elementary  Theory,  Curtis  Turbine.  —  Fig.  184  gives  a  dia- 
grammatic arrangement  of  the  blades  and  nozzles  in  the  first  stage 
of  a  two-stage  Curtis  turbine,  each  stage  consisting  of  one  set  of  nozzles 
and  two  moving  and  one  stationary  sets  of  blades.  The  object  of 
employing  a  number  of  stages  is  to  permit  of  a  low  peripheral  velocity 
without  reducing  the  efficiency.  Since  the  velocity  of  steam  varies  as 
the  square  root  of  the  kinetic  energy,  the  theoretical  stage  velocity  may 
be  determined  by  dividing  the  maximum  initial  velocity  by  the  square 
root  of  the  number  of  stages,  assuming  that  the  entire  velocity  of  the 
jet  is  abstracted  in  each  stage.  Thus  if  the  turbine  were  constructed 
with  four  stages  the  theoretical  stage  velocity  would  be  reduced  from 
say  3600  feet  per  second  to  1800  feet  per  second.  In  general,  in  order 
to  reduce  the  stage  velocity  to  V8  feet  per  second,  the  number  of  stages  n 
may  be  determined  from  the  equation 


n  = 


(77) 


in  which  V  ==  maximum  initial  velocity. 

Referring  to  Fig.  184:  the  steam  is  expanded  in  the  first  stage  from 
pressure  Pl  to  P2  and  issues  from  the  first  set  of  nozzles  with  absolute 
velocity  Vlt  striking  the  first  set  of  moving  blades  at  an  angle  a  with 
the  line  of  motion  of  the  wheel.  The  resultant  v1  of  Vl  and  the 


STEAM  TURBINES 


359 


peripheral  velocity  u,  is  the  velocity  of  the  steam  relative  to  the  vanes ; 
and  the  angle  /?  which  the  line  v1  makes  with  the  line  of  motion  of  the 
wheel  is  the  proper  entrance  angle  of  the  blades  for  the  first  set. 
Neglecting  friction  the  exit  angle  y  will  be  the  same  as  the  entrance 
angle  /?.  The  resultant  of  v2  the  exit  velocity  relative  to  the  blade,  and 
u,  the  peripheral  velocity,  is  V2,  the  absolute  exit  velocity. 


Nozzles 


FIG.  184.     Velocity  Diagram,  Curtis  Turbine. 

Since  the  second  set  of  blades  is  fixed  and  serves  as  a  means  of  changing 
the  direction  of  flow,  the  absolute  velocity  entering  them  is  V2.  The 
angle  d  formed  by  V2  and  the  center  line  of  the  stationary  blades  is  the 
proper  entrance  angle.  Neglecting  friction  the  absolute  exit  velocity 
will  be  F3  =  V2,  and  the  exit  angle  will  be  e  =  d.  The  steam  flowing 
from  the  stationary  blades  strikes  the  second  set  of  moving  blades  at 
an  angle  £  =  d  with  absolute  velocity  V3.  Combining  F3  with  the 


360  STEAM  POWER  PLANT  ENGINEERING 

peripheral  velocity  u  we  get  v3,  the  velocity  of  the  steam  relative  to  the 
second  set  of  moving  blades.  The  angle  6,  formed  by  v3  and  the  line 
of  motion  of  the  wheel,  is  the  proper  entrance  angle  for  the  second  set 
of  moving  blades.  The  resultant  of  v4  (=  v3)  and  u  is  V4,  the  absolute 
exit  velocity  for  the  first  stage.* 

In  the  second  stage  the  steam  is  expanded  from  pressure  P2  to  that 
in  the  condenser  and  acquires  initial  velocity  Va,  leaving  the  last  bucket 
with  residual  velocity  Vn-  The  theoretical  velocities  and  blade  angles 
for  this  stage  may  be  found  as  above. 

Example:  A  two-stage  Curtis  turbine  develops  800  horse  power  on  a 
steam  consumption  of  12.5  pounds  per  horse-power  hour,  steam  dry  and 
saturated.  Initial  gauge  pressure  135  pounds  per  square  inch  and  back 
pressure  2  pounds  per  square  inch  absolute.  Peripheral  velocity  600  feet 
per  second.  The  steam  expands  in  the  first  stage  from  150  pounds  to  20 
pounds  absolute,  and  in  the  second  stage  from  20  to  2  pounds  absolute. 
Angle  of  the  nozzles  with  the  plane  of  rotation,  20  degrees.  Compare  the 
performance  of  the  actual  turbine  with  its  theoretical  possibilities. 
Actual  turbine: 

Steam  consumed  per  hour  =  800  X  12.5  =  10,000  pounds. 

Steam  consumed  per  second  =  10,000  -r-  3,600  =  2.78  pounds. 

Horse  power  developed  per  pound  of  steam  flowing  per  second, 
800  -*-  2.78  =  288. 

Kinetic  energy  =  288  X  550  =  158,400  foot-pounds  per  second. 

Thermal  efficiency,  equation  (68), 

2545 
E<  -  12.5  (1191.1  -  94.4)  "  18'7  Per  cent' 

Heat  consumption,  B.T.U.  per  horse  power  per  minute, 
12.5  (1191.1  -  94.4)  = 
60 

Tfl  1  ft    7 

Efficiency  ratio  =  -=-'  =  ^^  =  0.739. 

/',  ,       _•)..) 

Ideal  turbine: 

Stage  velocities. 

The  theoretical  velocity  in  the  first  stage  in  expanding  from  a  pres- 
sure of  150  pounds  to  20  pounds  absolute  will  be 

Vl  =  224  VHl  -  H2 


=  224  V1191.1  -  1042.9 
=  2727  feet  per  second. 

*  In  the  actual  turbine  the  velocities  will  be  less  than  the  theoretical  on  account 
of  f  Fictional  resistances  in  the  nozzles  and  blades,  and  the  velocity  diagram  must  be 
modified  as  indicated  in  Fig.  172b  and  described  on  p. 


STEAM  TURBINES  361 

In  the  second  stage  the  steam  expands  from  20  pounds  to  2  pounds 
absolute.  _ 

Va  =  224  V#3  -H* 

=  224  V1042.9  -  914.8 
=  2530  feet  per  second. 

The  kinetic  energy  per  pound  of  steam  in  each  division  of  the  first 
stage  will  be: 

In  the  first  set  of  moving  blades, 

'  /^  -•-••• 


=  J_  (27272  -  16202)  =  74,722  foot-pounds  per  second. 

The  value  of  V2  is  conveniently  obtained  from  the  velocity  diagram. 
In  the  second  set  of  moving  blades, 


64.4 

1 
64.4 


(16202  -  9252)  =  27,448. 


Total  energy  in  first  stage  =  102,170. 

In  a  similar  manner  the  total  energy  in  the  second  stage  will  be  found 
to  be  93,700. 

Total  for  entire  turbine  =  195,870  foot-pounds  per  second. 
Theoretical  horse  power  per  pound  of  steam: 

195,870 
- 


Theoretical  steam  consumption  per  horse-power  hour, 

3600 

—-  =  10.1  pounds. 
dob 

Heat  consumption,  B.T.U.  per  horse  power  per  minute, 
10.1  (1191.1  -  94.4) 


60 
Thermal  efficiency  ratio, 


=  184. 


1191.1  -  914.8 
Er=~   1191.1-94.4  = 

*  In  the  actual  turbine  the  heat  contents  H2,  H3  and  H4  will  be  greater  than  that 
of  the  ideal  mechanism  on  account  of  fractional  losses.     See  equation  (76d). 


362 


STEAM   POWER   PLANT   ENGINEERING 
SUMMARY. 


Actual  Turbine. 

Perfect  Turbine. 

Horse  power  developed  per  pound  of  steam  

288 

356 

Steam  consumption,  pounds  per  H.P.  hour  
B  T  U   consumed  per  H  P   per  minute 

12.5 

288 

10.1 
184 

Thermal  efficiency   per  cent 

18.7 

25.3 

Efficiency  ratio,  per  cent  

73.9 

191.  The  Hamilton-Holzworth  Steam  Turbine.  —  Figs.  185  to  188 
give  a  general  view  and  some  of  the  details  of  the  Hamilton-Holzworth 
turbine,  which  belongs  to  the  compound  multi-stage  "  velocity  "  type. 
The  steam  flows  through  the  annular  space  between  rotor  and  stator 


CONDENSER 

OR 

EXHAUST 
PRESSURE 


BOILER         *•  P- 1 


FIG.  185.     Principles  of  Hamilton-Holzworth  Steam  Turbine. 

as  in  the  Parsons,  but  differs  from  the  latter  in  that  expansion  takes 
place  only  in  the  stationary  vanes.  The  rotor  consists  of  a  number 
of  steel  disks  of  varying  diameters  riveted  to  both  sides  of  steel  hubs 
and  fitted  at  the  periphery  with  drop-forged  vanes  as  shown  in  Fig. 
186.  A  tough  steel  ring  is  shrunk  on  the  outside  periphery  of  the  vanes 
as  indicated.  The  number  of  wheels  and  vanes  is  considerably  less 
than  in  the  Parsons  type.  The  stationary  vanes  are  fitted  in  steel 
disks  as  shown  in  Fig.  185,  and  the  latter  are  located  in  grooves  in 
the  turbine  casing.  The  vanes  have  a  varying  radial  height  increasing 
in  the  direction  in  which  the  steam  flows.  Sizes  under  750  kilowatts 


STEAM   TURBINES 


363 


have  but  one  turbine  casing,  but  larger  sizes  are  divided  into  two,  a 
high  and  a  low-pressure  turbine.  The  operation  is  as  follows:  Steam 
enters  the  high-pressure  casing  as  indicated  by  arrows  in  Fig.  185,  and 
passes  through  the  first  set  of  stationary  vanes,  extending  around  the 
whole  periphery,  which  direct  the  steam  at  the  proper  angle  against  the 
wheel  blades.  In  passing  through  the  stationary  vanes  the  steam  is 
expanded  down  to  the  pressure  in  the  first  stage  which  is  the  same  on 
both  sides  of  the  rotating  disk.  After  giving  up  part  of  its  energy,  the 
steam  expands  again  through  the  second  set  of  stationary  vanes  to  the 

pressure  in  the  second  stage, 
giving  up  energy  to  the  second 
set  of  moving  vanes.  This  pro- 
cess is  repeated  until  the  last 
stage  is  reached,  from  which 
the  steam  is  discharged  to  the 
condenser  in  the  simple  turbine, 
or  to  the  low-pressure  steam 
chest  in  the  compound  turbine. 
In  the  low-pressure  casing  the 
steam  is  distributed  in  the  same 
manner  as  in  the  high-pressure 
turbine.  The  diagram  in  the 
lower  part  of  Fig.  185  shows 
the  variation  in  steam  pressure 
and  velocity.  The  low-pressure 
front  head  is  provided  with  an 
auxiliary  nozzle  which  may  be 
supplied  with  live  steam  in  case 
of  overload.  The  builders  claim 
that  since  the  pressures  on  both 
sides  of  the  wheel  are  the  same, 

no  provision  is  necessary  for  axial  balancing  as  in  the  Parsons  standard 
turbine. 

Fig.  187  shows  a  sectional  view  of  the  bearing  and  stuffing  box 
for  the  shaft  at  the  point  where  it  passes  through  the  end  of  the  turbine 
casing.  The  shaft  is  turned  to  a  smaller  diameter  at  its  end  and  runs  in 
a  bushing  G  having  a  flange  bearing  against  the  inner  side  of  the 
pillow  block.  At  A  is  a  cylindrical  piece  attached  to  arid  rotating  with 
the  shaft.  This  piece  projects  into  an  annular  groove  in  the  piece  B, 
but  it  does  not  completely  fill  the  groove  and  a  circuitous  passage 
is  formed  through  which  the  steam  must  pass  before  reaching  the 
stuffing  box  C.  The  object  of  the  passage  is  to  provide  condensing 


DETAILS  OF  VAMEC. 
SECTION  OF   WHEEL. 

FIG.  186.     Details  of  Vanes,  Hamilton- 
Holz worth  Steam  Turbine. 


364 


STEAM  POWER  PLANT  ENGINEERING 


surface  so  the  steam  itself  will  not  reach  the  packing.     The  joint  at  the 
stuffing  box  is  thus  practically  water-sealed. 


FIG.  187.     Details  of  Bearing,  Hamilton-Holzworth  Turbine. 

To  prevent  the  oil  from  working  into  the  turbine  a  bushing  F  is 
attached  to  the  shaft  which  throws  off  the  oil  into  the  space  D  by  cen- 
trifugal force,  where  it  drips  down  through  a  channel  into  a  compart- 
ment in  the  pillow  block.  Any  water 
escaping  through  the  stuffing  box  is 
also  collected  in  the  same  compart- 
ment. The  bearing  is  oiled  by  a 
forced-oil  system,  the  oil  being  sup- 
plied to  the  bottom  of  the  bushing. 

Fig.  188  gives  a  diagrammatic  view 
of  the  governor  mechanism.  A  is  the 
friction  disk,  L  the  roller,  C  the  splined 
shaft  with  which  the  roller  turns  but 
upon  which  it  is  free  to  slide,  0  bevel 
gears  connecting  shaft  C  with  the 
throttle  valve,  and  E  a  worm  wheel 
driving  disk  A  by  means  of  a  worm 
shaft  F.  At  normal  speed  the  gov- 
ernor sleeve  is  in  mid  position  and 
roller  L  is  at  the  center  of  the  disk. 
If  the  turbine  speeds  up,  however,  the 
governor  sleeve  will  rise,  carrying  with 
it  the  right-hand  arm  of  lever  T,  which 
in  turn  will  push  the  roller  L  a  cor- 
responding distance  downward.  At 
the  same  time  the  cam  H  will  be 
thrown  to  the  right  by  contact  with  the  roller  R  and  by  means  of 
lever  U  will  move  the  disk  A  and  its  shaft  to  the  left,  bringing  it  in 


FIG.   188.        Governor  Mechanism, 
Hamilton-Holzworth  Turbine. 


STEAM  TURBINES  365 

contact  with  roller  L,  thus  imparting  a  rotary  motion  to  the  shaft  C 
and  closing  the  throttle  valve  until  the  turbine  assumes  normal  speed, 
when  the  several  parts  assume  the  first  position.  If  the  speed  is 
reduced  below  normal  the  operation  is  just  the  same  except  that  the 
various  motions  are  reversed. 

At  an  increase  in  speed  of  2}  per  cent  above  normal  steam  is  cut  off 
entirely. 

HamiUor^Holzwarth  Turbine:  Am.  Elecn.,  Oct.,  1904,  p.  549;  Eng.  Rec.,  Oct.  1, 
1904,  p.  405;  Power,  Dec.,  1907,  p.  878,  Nov.,  1904,  p.  659;  Machinery,  Nov.,  1904, 
p.  134;  Engr.  U.S.,  Oct.  1,  1904,  p.  690. 

192.  Westinghouse-Parsons  Steam  Turbine. — Fig.  189  shows  a  section 
through  a  Westinghouse-Parsons  multi-stage  reaction  turbine.  In  this 
type  no  nozzles  are  employed  and  expansion  of  the  steam  is  effected 
by  a  series  of  stationary  and  movable  blades.  The  rotor  is  a  steel 
barrel  or  drum  divided  into  three  sections  of  varying  diameter,  upon 
the  periphery  of  which  bronze  blades  are  radially  inserted  in  dove- 
tailed grooves.  The  adoption  of  three  sections  of  varying  diameter 
has  no  bearing  on  the  design  of  this  machine  but  is  merely  for 
mechanical  convenience.  The  blades  increase  in  length  and  cross 
section  from  the  high-pressure  to  the  low-pressure  end  of  each  section. 
The  stator  is  of  cast  iron  and  its  inner  surface  is  studded  with  rows  of 
blades  projecting  radially  inward  and  conforming  in  size  with  the  adjoin- 
ing blades  of  the  rotor.  The  relative  positions  of  the  blades  in  the  rotor 
and  stator  are  shown  in  Fig.  190.  The  operation  of  the  turbine  is  as 
follows:  Steam  enters  at  S,  Fig.  189,  through  poppet  valve  V,  which  is 
actuated  by  the  governor  shown  in  detail  in  Fig.  191,  and  flows  through 
the  annular  space  between  rotor  and  stator  to  the  exhaust  opening 
at  B.  The  entire  expansion  is  carried  out  within  this  annular  compart- 
ment and  resembles  in  effect  a  simple  divergent  nozzle  with  the  excep- 
tion that  the  dynamic  relationship  of  jet  and  vane  is  such  as  to  secure 
a  comparatively  low  velocity  from  inlet  to  exhaust.  The  velocity 
varies  from  150  feet  per  second  at  the  high-pressure  end  to  about  600 
feet  per  second  as  a  maximum  at  the  low-pressure  end.  The  action 
of  the  steam  on  the  blades  is  illustrated  in  Fig.  190.  The  steam  strikes 
the  first  set  of  stationary  blades  as  at  P  with  initial  velocity  of  about 
150  feet  per  second  and  is  deflected  against  the  moving  blades  immedi- 
ately adjoining.  In  passing  from  P  to  Pl  the  steam  is  partly  expanded 
and  gives  up  a  portion  of  its  energy  to  the  moving  blades.  The  steam 
is  deflected  from  Pl  to  Pn  and  thus  has  a  reactive  effect  on  the  moving 
blades  in  addition  to  the  impulse  imparted  at  Pr  The  total  torque 
produced  at  the  shaft  in  element  A  is  therefore  due  to  impulse  from 


366 


STEAM  POWER  PLANT  ENGINEERING 


STEAM  TURBINES 


367 


1  and  reaction  from  2.  This  process  is  repeated  in  each  element  of 
the  turbine,  the  steam  expanding  as  it  flows  from  element  to  element 
in  its  passage  to  the  condenser.  The  angular  velocity  of  the  rotor 


i< 


Stationary 


uufcccc 

Njpm 


Stationary 


Blades 


ving  Blades 


Blades 


J)  J)  J)  J);j)  J)  J) 


Blades 


FIG.  190.     Flow  of  Steam  in  Parsons  Turbine. 

varies  from  3600  r.p.m.  in  a  400-kilowatt  unit  to  750  r.p.m.  in  the  7500- 
kilowatt  size.  Opposed  to  the  three  sets  of  blades  the  spindle  also  carries 
three  rotating  balance  pistons  P,  P,  Fig.  189,  each  of  such  diameter  as 


FIG.  191.     Governor  Mechanism,  Westinghouse-Parsons  Turbine. 

to  exactly  balance,  through  passage  E,  the  axial  thrust  of  the  steam 
against  its  corresponding  drum  of  blades. 

Steam  enters  the  turbine  intermittently  as  shown  in  Fig.  192,  which 
represents  indicator  cards  from  a  1250-kilowatt  turbine  at  various  loads. 


368 


STEAM  POWER  PLANT  ENGINEERING 


At  light  load  the  valve  opens  for  a  very  short  period  and  remains  closed 
during  the  greater  part  of  the  interval.  As  the  load  increases,  the  period 
lengthens  until  finally,  at  about  full  load,  the  valve  does  not  reach  its 
seat  at  all,  and  continuous  pressure  is  obtained  in  the  high-pressure  end 
of  the  turbine. 

The  intermittent  admission  of  steam  is  produced  and  controlled  as 
follows:  Lever  T,  Fig.  191,  is  given  a  reciprocating  motion  by  an  eccentric 
actuated  by  a  worm  and  worm  wheel  on  the  main  shaft.  This  motion 
is  transmitted  through  lever  H  (with  fixed  fulcrum  B]  to  lever  A  (with 
floating  fulcrum  D)  and  finally  to  pilot  valve  G.  This  reciprocating 
pilot  valve  admits  puffs  of  steam  from  pipe  0  to  the  under  side  of 
piston  My  the  rod  R  of  which  is  attached  to  the  admission  valve  V  in 
Fig.  189.  A  spiral  spring  holds  piston  M  in  its  lowest  position  until 


FIG.  192.     Indicator  Cards  Showing  Initial  Pressure  in  a  Westinghouse-Parsons 

Steam  Turbine. 

steam  admitted  by  the  pilot  overcomes  the  spring  tension  and  lifts  the 
main  valve  from  its  seat,  thereby  permitting  steam  to  enter  the  turbine. 
The  fulcrum  D  of  lever  A  is  raised  and  lowered  by  the  governor  and 
therefore  the  pilot  valve  is  controlled  both  by  the  motion  of  the  eccen- 
tric and  the  motion  of  the  governor.  The  eccentric  keeps  the  pilot 
valve,  and  hence  the  main  throttle,  in  constant  oscillation,  while  the 
movement  of  the  governor  changes  the  limits  of  this  motion. 

If  an  overload  is  sufficiently  great  to  cause  the  governor  balls  to  drop 
to  their  lowest  position,  the  auxiliary  or  secondary  valve  Vs,  Fig.  189, 
begins  to  open  and  admits  high-pressure  steam  to  the  later  stage  where 
the  working  steam  areas  are  greater,  thus  increasing  in  proportion  the 
total  power  of  the  turbine.  The  operation  of  this  valve  is  the  same  as 
the  main  admission  valve  and  is  controlled  by  the  governor.  Fig.  193 
shows  the  details  of  this  mechanism.  The  speed  varies  about  2  per 
cent  from  no  load  to  full  load. 


STEAM  TURBINES 


369 


In  the  smaller  size  machines  running  above  1200  r.p.m.,  flexible  bear- 
ings are  employed  to  absorb  the  vibration  incident  to  the  critical  velo- 
city. They  consist  of  a  nest  of  loosely  fitting  concentric  bronze  sleeves 
with  sufficient  clearance  between  them  to  insure  the  formation  of  a 
film  of  oil.  In  the  larger  machines  running  below  1200  r.p.m.  a  split 
self-aligning  bearing  is  used  instead  of  the  -flexible  bearing.  The  ends 
of  the  casing  are  fitted  with  water-sealed  glands  of  special  design  to 
prevent  the  escape  of  steam  or  inflow  of  air  at  the  point  of  entry  of  the 
shaft.  The  water  used  for  sealing  them  is  small  in  quantity  and  may 
be  returned  to  the  feed-water  system. 


FIG.  193.     By-Pass  Valve,  Westinghouse-Parsons  Turbine. 

Double-flow  Type.  —  In  reaction  turbines  of  .  the  single-flow  type, 
as  illustrated  in  Fig.  189,  the  high-pressure  portion  dealing  with  the 
high-pressure  incoming  steam  is  the  least  efficient.  This  is  due  to 
the  fact  that  the  blade  lengths  are  approximately  proportional  to  the 
specific  Volume  of  the  steam,  and  consequently  the  initial  expansion 
in  the  turbine  requires  blade  passages  of  very  small  dimensions.  This 
results  in  greater  leakage  past  the  tips  of  the  blades  than  in  the  low- 
pressure  elements  where  the  blades  are  long.  Again,  in  the  single-flow 
type  the  high-pressure  balance  piston  occupies  fully  one-half  of  the 
total  balance  piston  length  of  the  shaft,  while  the  low-pressure  piston 
is  2^  times  the  high-pressure  diameter,  so  that  balance  pistons  occupy 
a  large  portion  of  the  total  bulk  of  the  machine.  By  making  the  high- 
pressure  element  of  the  impulse  type  and  by  arranging  the  low-pressure 
reaction  elements  on  either  side  as  illustrated  in  Fig.  193b  the  efficiency 
may  be  increased  and  the  bulk  of  the  turbine  may  be  greatly  decreased. 


370 


STEAM  POWER  PLANT  ENGINEERING 


There  are  two  rows  of  moving  buckets  upon  the  impulse  wheel  with  an 
intermediate  set  of  reversing  blades,  the  operation  being  practically 
the  same  as  in  the  first  stage  of  a  Curtis  turbine.  The  drop  in  pressure 
in  the  nozzles  is  such  that  approximately  20  per  cent  of  the  total  energy 
developed  is  absorbed  by  this  impulse  element.  After  leaving  the  im- 
pulse element  the  steam  divides,  one  portion  passing  directly  to  the 
low-pressure  blading  at  the  left,  while  the  rest  passes  through  the  hollow 
shell  of  the  rotor  to  the  similar  pressure  blades  upon  the  right.  As 
these  sections  are  equal  and  symmetrical  they  counterbalance  each 
other,  and  the  balance  or  "  dummy  "  pistons  may  be  dispensed  with. 
The  advantages  of  the  double-flow  type  over  a  single-flow  unit  of  equal 
capacity  are:  (1)  reduction  of  nearly  50  per  cent  in  the  shaft  span 


IMF: 


Plan  View 


Side  View  End  View 

FIG.  193a.     Method  of  Fastening  Blades  in  Westinghouse-Parsons  Turbines. 

between  bearings;  (2)  the  diameters  of  the  casing  and  rotating  part 
are  more  uniform,  thus  tending  to  greater  rigidity;  (3)  a  reduction  of 
about  70  per  cent  in  the  bulk  of  the  main  parts  of  the  machine,  and  (4) 
internal  stresses  due  to  high-pressure  and  high-temperature  steam  are 
avoided  by  isolating  the  incoming  steam,  without  separate  nozzle 
chambers.  Westinghouse-Parsons  turbines  are  made  in  a  number  of 
sizes,  varying  from  400  kw.  to  15,000  kw.  In  Europe,  however, 
Parsons  turbines  are  made  as  small  as  20  kw.  and  as  large  as  25,000 
H.P.  In  sizes  up  to  3500  kw.  the  single-flow  turbine  has  established 
itself  as  the  most  suitable  prime  mover,  but  for  larger  sizes  the  double- 
flow  is  given  preference.  The  double-flow  turbine  is  admirably  adapted 
to  low-pressure  work.  Fig.  194b  shows  a  section  through  a  Westing- 
house-Parsons double-flow  low-pressure  turbine.  For  results  of  tests 
of  Parsons  and  Westinghouse-Parsons  turbines  see  Table  48. 

Double-flow  Turbine:  Power  &  Engr.,  March  16,  1909,  Aug.,  1908,  p.  471 ;  Eng. 
Rec.,  May  30,  1908,  p.  693;  Elec.  Review,  June  26,  1908,  p.  1089. 

10,000-kw.  Westinghouse-Parsons  Double-flow  Turbo  Generator  for  the  Metropolitan 
Street  Railway  Company,  Kansas  City,  Kansas:  Power  &  Engr.,  May,  17,  1910,  p.  890. 


STEAM  TURBINES 


371 


372 


STEAM  POWER  PLANT  ENGINEERING 


192a.   Allis-Chalmers   Steam   Turbine.  —  Fig.   193c  shows  a  section 
through  an  Allis-Chalmers  standard  steam  turbine,  which  is  of   the 

Parsons  type  but  differs 
from  the  original  Parsons 
machine  and  the  Westing- 
house-Parsons  construc- 
tion principally  in  manu- 
facturing details.  In  the 
older  Parsons  type,  three 
balancepistons  are  placed 
at  the  high-pressure  end. 
In  the  Allis-Chalmers  de- 
sign, the  larger  piston  is 
placed  at  the  low-pressure 
end  of  the  rotor,  behind 
the  last  row  of  blades, 
the  other  two  remaining 
at  the  high-pressure  end. 
This  construction  per- 
mits of  a  smaller  balance 
piston  and  allows  asmaller 
working  clearance  in  the 
7/IH  B  •  IgjlA  _§  high-pressure  and  inter- 

,/ IjHi || •  ~  Bw  J a_  a      mediate  cylinders.  In  the 

Allis-Chalmers  turbine 
the  roots  of  the  blades  are 
dovetailed  and  fitted  into 
a  foundation  ring,  and  the 
tips  are  encased  in  a  chan- 
nel-shaped shroud  ring, 
thereby  insuring  a  rigid 
and  positively  spaced  con- 
struction. The  governor 
is  of  the  Parsons  type,  ex- 
cept that  the  main  valve 
and  pilot  valve  are  actu- 
ated by  hydraulic  instead 
of  steam  pressure.  The 
bearings  are  of  the  self- 
adjusting  ball  and  socket  pattern  and  are  kept  "  floating  in  oil  "  by 
a  small  pump  geared  to  the  turbine  shaft.  The  oil  is  passed  through 
a  tubular  cooler  with  water  circulation  after  it  leaves  the  bearings  and 
is  used  over  and  over  again. 


STEAM  TURBINES 


373 


193.  Elementary  Theory,  Parsons  Turbine.  —  Fig.  194  gives  a  dia- 
grammatic arrangement  of  fixed  and  stationary  blades  in  the  first 
stages  of  a  multi-stage  ideal  reaction  turbine.  The  steam  enters  the 
stationary  blades  at  practically  zero  velocity  and  is  there  partially 
expanded  and  impinges  against  the  movable  blades  at  velocity  Fu 
part  of  the  energy  of  the  steam  being  thus  absorbed.  In  passing 
through  the  movable  blades  the  steam  is  still  further  expanded  and 
leaves  at  an  absolute  velocity  F,,  exerting  an  additional  pressure  on 
the  blades  from  the  reaction.  The  steam  enters  the  second  set  of 
stationary  blades  with  velocity  F2  and  is  still  further  expanded  to 
velocity  F3,  and  so  on. 

'  P  =150 


is 


FIG.  194.     Velocity  Diagram.    Westingho use-Parsons  Turbine. 

The  energy  imparted  to  the  steam  in  the  first  set  of  stationary  blades 

Ei  =  -fg-      V'-  (78) 

Vl  =  absolute  velocity  of  the  steam  leaving  the  blades. 

The  energy  imparted  to  the  steam  in  the  first  set  of  moving  blades  is 

E2  =  ~-  (v.*  -  v').  (79) 

i\  =  relative  velocity  of  the  steam  entering  the  moving  blades. 
v2  =  relative  velocity  of  the  steam  leaving  the  moving  blades. 
The  total  energy  acquired  by  the  steam  in  the  first  stage  is 


E2. 


The  energy  converted  into  work  in  this  stage  is 
E  =  El  +  E2  -  WVjL 


W 


(80) 
(81) 


F2  =  absolute  velocity  of  the  steam  leaving  the  moving  blades. 
Each  stage  may  be  analyzed  in  a  similar  manner. 


374  STEAM  POWER  PLANT  ENGINEERING 

Example:  A  Westinghouse-Parsons  turbine  develops  1000  horse 
power  on  a  steam  consumption  of  12  pounds  of  steam  per  horse-power 
hour.  Initial  steam  pressure  150  pounds  per  square  inch  absolute; 
back  pressure  1  pound  per  square  inch  absolute;  drop  in  pressure  in 
each  set  of  fixed  and  moving  blades  15  pounds  per  square  inch; 
peripheral  velocity  300  feet  per  second;  «1  =  a2  =  30  degrees.  Com- 
pare the  performance  of  the  actual  and  ideal  turbine. 

Actual  turbine: 

Steam  consumed  per  hour, 

1000  X  12  =  12,000  pounds. 
Steam  consumed  per  second, 

12,000  -J-  3,600  =  3.33  pounds. 
Horse  power  developed  per  pound  of  steam  flowing  per  second, 

1000  ^  3.33  =  300. 
Kinetic  energy  per  pound  of  steam, 

300  X  550  =  165,000  foot-pounds  per  second. 
Thermal  efficiency, 


.  12  (11-70) 

Heat  consumption,  B.T.U.  per  horse-power  hour  per  minute, 
12(1191.2-70) 

-   =    ZZ1!. 

60 
Efficiency  ratio, 

..:  |  =  m.  67.5  percent.;;          '         \ 

Ideal  turbine: 

The  velocity  imparted  to  the  steam  in  the  first  set  of  stationary 
blades  due  to  the  drop  from  150  to  135  pounds  per  square  inch  is 


V,  =224  VH,- 


=  224  V1191.2- 1182.4 
=  662  feet  per  second. 

Lay  off  the  value  of  Vl  in  direction  and  amount  and  combine  with  u, 
the  peripheral  velocity,  Fig.  194.  The  resultant  is  vlf  the  velocity 
of  the  steam  relative  to  the  blades.  The  angle  between  vl  and  the  line 
of  motion  of  the  wheel  will  be  the  angle  with  the  blade  at  entrance. 

From  the  velocity  diagram, 

v,  =  429. 


STEAM  TURBINES  375 

E2,  the  energy  given  up  by  one  pound  of  steam  in  expanding  from 
135  to  120  pounds,  is 

E2=  778  (#2-  #3) 

=  778  (1182.4-1172.8) 

=  7468  foot-pounds  per  second. 

Substitute        vl  =  429  and  E2  =  7468  in  equation  (79), 

'    7468  =  ^(^-429*),         i;;:  '::•-       ''    ' 

v2  =  816  feet  per  second. 

The  resultant  of  v2  and  u  is  V2,  the  absolute  velocity  of  the  steam 
leaving  the  moving  blades  of  the  first  stage.  From  the  diagram, 

F2  =  573  feet  per  second. 

The  energy  converted  into  work  in  the  first  stage  is  determined  by 
substituting  the  proper  values  in  equation  (81),  thus: 

E  =  (6622  +  8162  -  4292  -573*)  — 

64.4 

=    9200  foot-pounds  per  second. 

The  various  stages  may  be  analyzed  in  a  similar  manner. 

The  theoretical  output  of  the  entire  turbine  per  pound  of  steam  will 
be  that  corresponding  to  adiabatic  expansion  from  a  pressure  of  150  to 
1  pound  absolute. 

E=778(Hl-  Hn) 
=  778(1191.2-877) 
=  244,447  foot-pounds  per  second. 

Horse  power  per  pound  of  steam, 


Steam  consumption  per  horse-power  hour, 
3600 


445 

Thermal  efficiency, 

Er 


=  8.1  pounds. 

=  1191.2-877 

1191.2-70 

=  28  per  cent. 


376 


STEAM  POWER  PLANT  ENGINEERING 


194.  Low-pressure  and  Mixed-pressure  Turbines.  —  A  promising 
field  for  the  steam  turbine  is  in  its  application  as  a  secondary  or  low- 
pressure  unit  in  connection  with  non-condensing  or  condensing  engines, 
or,  combined  with  a  regenerator,  in  connection  with  engines  using  steam 
intermittently.  Numerous  examples  may  be  cited  showing  great  gains 
in  both  capacity  and  economy  in  existing  power  plants  involving  the 


FIG.  194a.     Low-pressure  Turbine  Installation  at  the  59th  Street  Station  of  the 
Interborough  Rapid  Transit  Company,  New  York. 

abandonment  of  but  a  negligible  part  of  the  equipment  and  accom- 
plishing this  result  with  a  minimum  additional  investment.  The  most 
notable  installation  (June,  1910)  of  low-pressure  turbines  to  con- 
densing reciprocating  engines  is  at  the  59th  Street  Station  of  the 
Interborough  Rapid  Transit  Co.,  New  York.  Three  of  the  nine  7500- kw. 
Manhattan-type  compound  Corliss  engines  have  been  equipped  with 
Curtis  three-stage,  low-pressure  turbo-generators  of  equal  capacity,  and 
provision  is  made  for  the  installation  of  six  additional  units.  The  low- 
pressure  turbine  is  installed  between  the  exhaust  of  the  low-pressure 
cylinders  and  the  condenser  as  shown  in  Fig.  194a.  Running  with  the 


STEAM  TURBINES 


377 


378 


STEAM  POWER  PLANT  ENGINEERING 


engine  the  low-pressure  turbine  generator  carries  a  variable  load  without 
governor  regulation.  The  turbine  generator  takes  care  of  the  speed 
by  automatically  taking  such  a  load  as  will  keep  the  frequency  in  unison 


Engine  Load  K.W. 
3000         I         4000  5000 


FIG.  194c.     Performance  of  7500-Kw.  Engine  at  59th  Street  Station  of  Interborougli 
Rapid  Transit  Company,  New  York,  with  Varying  Receiver  Pressure. 


A:-  Constant  Nozzle  Pressure 


B=-  Variable  Nozzle  Pressure 

All  Nozzles  Open 
Corrected  for  Moisture  in 
Turbine  Steam 


7000    8000 


9000 


10000   11000   12000   13000 
Unit  Load-JC.W. 


14000   15000   16000 


FIG.  194d.  Comparison  of  Economy  Curves:  7500-Kw.  High-pressure  Turbine,  7500-Kw. 
Engine  and  Combined  Engine  and  Low-pressure  Turbine  at  the  59th  Street  Station 
of  the  Interborough  Rapid  Transit  Company,  New  York. 

with  that  of  the  engine-driven  generator.  The  turbine  is  equipped 
with  the  usual  emergency  speed  limit  attachment  for  cutting  off  the 
steam  supply  should  the  speed  exceed  a  predetermined  limit.  The 


STEAM  TURBINES 


379 


performance  of  one  set  of  engines,  a  high-pressure  turbine  of  the  equiva- 
lent total  capacity,  and  that  of  the  combined  engine  and  low-pressure 
turbine,  are  illustrated  in  Fig.  194d.  The  conclusions  drawn  from  an 
exhaustive  series  of  tests  at  this  station  are  that  the  addition  of  low- 
pressure  turbines  effected: 

a.  An  increase  of  100  per  cent  in  maximum  capacity  of  plant. 

b.  An  increase  of  146  per  cent  in  economic  capacity  of  plant. 

c.  A  saving  of  approximately  85  per  cent  of  the  condensed  steam  for 
return  to  the  boiler. 

d.  An  average  improvement  in  economy  of  13  per  cent  over  the 
best  high-pressure  turbine  results. 

e.  An  average  improvement  in  economy  of  25  per  cent  (between  the 


FIG.  195.     Rateau  Low-Pressure  Steam  Turbine  Installation. 

limits  of  7000  kw.  and  15,000  kw.)  over  the  results  obtained  by  the 
engine  units  alone. 

/.  An  average  unit  thermal  efficiency  of  20.6  per  cent  between  the 
limits  of  6500  kw.  and  15,500  kw. 

Low-pressure  turbines  are  frequently  installed  in  connection  with 
regenerative  accumulators,  to  rolling-mill  engines,  steam  hammers,  and 
other  appliances  using  steam  intermittently,  and  have  proved  to  be 
paying  investments.  A  typical  installation  of  this  character  is  to  be 
found  at  the  South  Chicago  Division  of  the  International  Harvester 
Company.  The  front  elevation  of  the  turbine  and  regenerator  installa- 
tion is  shown  in  Fig.  195  and  the  general  arrangement  of  the  regenerator 
is  shown  in  Fig.  196.  The  regenerative  accumulator  is  intended  to 
regulate  the  intermittent  flow  of  steam  before  it  passes  to  the  turbine. 


380 


STEAM  POWER  PLANT  ENGINEERING 


The  steam  collects  and  is  condensed  as  it  enters  the  apparatus  and  is 
again  vaporized  during  the  time  when  the  exhaust  of  the  engines  dimin- 
ishes or  ceases. 

The  regenerator  consists  of  a  cylindrical  boiler-steel  shell  divided 
into  two  similar  chambers  by  a  central  horizontal  diaphragm.  In  each 
compartment  are  a  number  of  elliptical  tubes  A,  each  of  which  is  per- 
forated with  a  number  of  f-inch  holes.  The  spaces  surrounding  the 
tubes  and,  under  certain  conditions,  the  tubes  themselves  are  filled 
with  water  to  a  height  of  about  four  inches  above  the  top  of  the  upper 
tubes.  Baffle  plate  B  serves  to  separate  the  entrained  moisture  from 
the  steam.  The  operation  is  as  follows:  Exhaust  steam  enters  the 
apparatus  at  N,  passes  to  the  interior  of  the  elliptical  tubes,  and  escapes 


FIG.  196.     Rateau  Regenerator  Accumulator. 

into  the  steam  space  through  the  perforations  and  thence  to  the  turbine. 
When  the  supply  of  steam  from  the  main  engine  ceases,  the  pressure  in 
the  regenerator  decreases,  the  water  liberates  part  of  the  heat  it  has 
absorbed  and  a  uniform  flow  of  low-pressure  steam  is  given  off.  The 
continued  demand  of  the  turbine  reduces  the  pressure  in  the  accumula- 
tor and  causes  the  steam  still  retained  in  the  tubes  to  escape,  thereby 
maintaining  the  circulation  of  the  water  (indicated  by  arrowheads)  and 
facilitating  the  liberation  of  steam.  Suitable  valves  regulate  the  limits 
of  pressure  in  the  accumulator  and  prevent  the  return  of  water  to  the 
main  engine.  Low-pressure  turbines  develop  one  electrical  horse- 
power hour  on  a  steam  consumption  of  about  30  pounds  with  initial 
pressure  of  15  pounds  absolute  and  a  back  pressure  of  1.5  pounds 
absolute.  Fig.  197  gives  the  performance  of  the  500-kilowatt  Rateau 
turbine  at  the  International  Harvester  Works,  South  Chicago,  1., 


STEAM  TURBINES 


381 


average   initial   pressure   of    16   pounds   absolute,    condenser  pressure 
1.5  pounds  absolute. 

Low-pressure  turbines  equipped  with  special  expanding  nozzles,  or 
the  equivalent,  to  receive  steam  at  high  pressure  direct  from  the  boilers 
are  known  as  mixed  pressure  turbines.  With  this  construction  the 
full  power  of  the  turbine  can  be  developed  with  (1)  all  low-pressure 
steam,  (2)  all  high-pressure  steam,  (3)  any  proportion  of  high  and  low 


Reinforced  Cinder  Concrete 


FIG.  196a.     Typical  Double-deck  Installation,  Fort  Wayne  and  Wabash  Valley 
Traction  Company,  Spy  Run  Station. 

pressure  steam.  In  the  Curtis  mixed-pressure  turbine  this  transition 
from  all  low  pressure  to  all  high  pressure,  through  all  the  conditions 
intermediate  between  these  extremes,  is  provided  for  automatically 
by  the  turbine  governor;  a  deficiency  of  low-pressure  steam  causes  the 
high-pressure  nozzles  to  open  automatically.  With  this  arrangement 
it  is  not  necessary  for  purposes  of  economy  to  proportion  exactly  the 
low-pressure  turbine  to  the  amount  of  exhaust  steam  available,  but 
within  limits  it  may  be  made  as  large  as  the  load  demands. 

Low-pressure  Turbines:  Power  &  Engr.,  July  6,  1909,  p.  1,  Nov  30,  1909,  p.  905; 
Prac-  Engr.  U.  S.,  Mar.  1,  1909,  p.  169;  Eng.  Mag.,  Apr.  and  May,  1907;  Iron  Age, 
Jan.  7,  1909. 


382 


STEAM   POWER  PLANT  ENGINEERING 


195.  Advantages  of  the  Steam  Turbine.  —  The  principal  advantages 
of  the  steam  turbine  are  (1)  simplicity;  (2)  economy  of  space  and 
foundation;  (3)  absence  of  oil  in  condensed  steam;  (4)  freedom  from 
vibration;  (5)  uniform  angular  velocity;  (6)  large  overload  capacity; 
and  (7)  high  efficiencies  for  large  variations  in  load.  The  reciprocating 
engine  is  well  adapted  for  pumping  stations,  direct-current  generators, 
compressor  plants,  hoisting  engines,  and  the  like,  requiring  low  angular 
velocity,  but  its  place  is  being  rapidly  taken  by  the  steam  turbine  for 
alternating-current  dynamos,  centrifugal  pumps  and  blowers,  requiring 
high  angular  velocity. 


:±HmHH±EHH:HH 

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-    CURVE  OF  STEAM  CONSUMPTION    - 
RATEAU  STEAM  TURBINE 

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300 


400 


500  600 

H.P.  at  Switchboard 

FIG.  197. 


700 


800 


900 


196.  Simplicity.  —  Although  composed  of  a  large  number  of  parts 
as  compared  with  a  reciprocating  engine  of  the  same  capacity,  there  are 
few  moving  parts  and  rubbing  surfaces.     The  only  contact  between 
rotor  and  stator  is  in  the  main  bearings,  and  the  problem  of  lubrication 
is  therefore  a  simple  one.     The  absence  of  pistons,  stuffing  boxes,  dash 
pots,  etc.,  reduces  the  cost  of  maintenance  and  attendance  to  a  minimum 
and  limits  the  possibility  of  leakage. 

197.  Economy  of  Space  and  Foundation.  —  Fig.  198  (J.  R.  Bibbins, 
Power,  January,  1905)  shows  the  relative  floor  space  required  for  different 
types   of  slow-speed  reciprocating  engines  and  Westinghouse-Parsons 
steam  turbines.     The  floor  space  required  by  the  Curtis  turbine  is  con- 
siderably less  than  for  the  Parsons  type,  as  it  is  vertical  with  the  gener- 
ator mounted  above  the  turbine.     Vertical   three-cylinder   compound 


STEAM  TURBINES 


383 


Corliss  engines  of  the  New  York  Edison  type  require  the  least  floor 
space  of  any  large  slow-speed  reciprocating  engines,  but  take  up  about 
twice  the  space  of  a  Parsons  turbine  installation  of  the  same  size. 
With  high-speed  engines  of  the  Willans  central-valve  type  the  com- 
parative economy  in  space  is  less  marked.  The  curves  refer  to  the 
space  occupied  by  engine  and  generator  alone,  whereas  in  a  modern 
turbine  installation  with  surface  condenser  the  condenser  equipment 
may  occupy  as  much  space  as,  if  not  more  than,  the  turbine  proper, 
and  considering  the  small  space  taken  up  by  the  barometric  condenser, 
such  as  is  used  in  connection  with  the  New  York  Edison  engines,  the 
economy  is  not  so  evident*  In  recent  turbine  installations  the  baro- 


i.o 


REPRESENTATIVE  TYPES  OF  PRIME  MOVERS 

Based  on  Overall  Dimensions 

of  Generating  Unit 


COMPARATIVE  FLOOR  SPACE  FOR 


1000 


2000 


3000     400.0     5000 

FIG.  198. 


6000 


7000 


8000 


metric  condenser  is  finding  much  favor,  and  in  such  instances  the  curves 
may  be  taken  to  indicate  the  relative  floor  space  for  the  entire  equip- 
ment of  prime  mover  and  auxiliaries. 

The  weight  of  the  steam  turbine  is  very  small  compared  with  a  recip- 
rocating engine  of  the  same  horse  power.  The  New  York  Edison  engines 
and  generators  weigh  more  than  eight  times  as  much  as  a  turbine 
installation  of  equal  capacity.  The  turbine  for  this  reason,  and  also 
because  of  the  total  absence  of  vibration,  requires  a  relatively  light 
foundation.  In  many  instances  the  foundation  consists  of  steel 
beams  with  concrete  arches  sprung  between  them  resting  upon  the 
floor,  and  the  basement  underneath  may  be  used  for  the  condenser 
instead  of  the  massive  foundation  required  for  the  reciprocating  engine. 


384  STEAM  POWER  PLANT  ENGINEERING 

198.  Absence  of  Oil  in  Condensed  Steam.  —  As  the  steam  turbine 
requires  no  internal  lubrication,  oil  does  not  come  in  contact  with  the 
steam,  and  the  condensed  steam  from  the  surface  condensers  is  available 
for  boiler-feeding  purposes  without  purification.     In  many  cases  the 
re-use  of  condensed  steam  effects  a  large  saving  in  cost  of  feed  water 
and  in  expense  for  maintenance  and  cleaning  of  boilers.     The  amount 
of  entrained  air  is  reduced  to  a  minimum  and  consequently  the  work 
of  the  air  pumps  lessened. 

199.  Regulation.  —  The   variable  pressure   at   the   crank   pin   of   a 
reciprocating  engine  necessitates  the  use  of  a  heavy  fly  wheel  to  keep 
the  instantaneous  angular  fluctuation  within  practical  limits.     In  the 
steam  turbine  the  motion  is  purely  rotary  and  a  fly  wheel  is  not  neces- 
sary.    In   the   former   there   are   always   instantaneous   variations   in 
velocity   during  each  revolution,   even  with   constant  load,   while  in 
the  latter  the  speed  is  practically  constant.      A  number  of  published 
tests  of  Parsons  and  Curtis  turbines  show  an  average  fluctuation  of 
2  per  cent  from  no  load  to  full  load  and  3  per  cent  from  no  load 
to   100  per  cent  overload.     Although  closer  regulation   than   this   is 
possible,    it   is    not    deemed    necessary,   particularly   in    alternating- 
current    work    where   a    comparatively    wide    range   is    desirable   for 
parallel  operation. 

200.  Overload   Capacity.  —  A  particular   advantage  of  the  turbine 
over  the  reciprocating  engine  lies  in  its  greater  overload  capacity  and 
higher  economy  at  overloads.     The  maximum  economy  of  the  average 
reciprocating   engine    lies    between   0.75    and   full   load,   whereas   the 
turbine  reaches  its  maximum  at  about  25  per  cent  overload.     Thus 
a  single  turbine   unit   may   economically   take  the  place  of  two   or 
more   reciprocating   units  for  a   variable   load.      A   turbine   may   be 
readily  operated  at  100  per  cent  overload,  while  the  ordinary  engine 
reaches  its  maximum  capacity   at   about  50  per  cent  overload.      In 
central  lighting  and  power  stations  where  there  are  one  or  more  sharp 
peak  loads  of  short   duration,  this  extreme  overload  capacity  is   of 
marked  importance. 

201.  Efficiency  and  Economy.  —  As  far  as  steam  consumption  is 
concerned  there  is  practically  no  difference  between  the  performance 
of  a  high-grade  piston  engine  and  that  of  a  first-class  turbine  for  sizes 
under  2000  kw.,  the  choice  depending  more  upon  rotative  speed,  over- 
load capacity  and  space  requirements  than  upon  the  heat  economy. 
For  sizes  over  2000  kw.   the  fuel  consumption  lies  in  favor  of  the 
turbine.     A  comparison  of  Fig.  148,  showing  typical  economy  curves  of 
high-speed  single- valve,  non-condensing  engines,  and  of  Fig.  198c,  show- 
ing similar  curves   for  small   non-condensing  turbines,    is   somewhat 


STEAM  TURBINES 


385 


in  favor  of  the  piston  engine,  though  the  difference  is  small;  whereas  a 
comparison  of  the  turbine  and  engine  curves  in  Fig.  194d,  showing  the 
performance  of  very  large  units,  is  decidedly  in  favor  of  the  turbine. 


£H 

w 

|3200 

|  2800 
GO 
2400 
55 
2000 
50 
IGOO 
45 

1200 
40     g 

35     * 
an    ^ 

^ 

^^x 

^ 

^ 

-^ 

y 

^ 

** 

,-C 

\\ 

^ 

^ 

fr 

& 

V 

^* 

V 

\ 

Steam  Pressure    175  # 
Back  Pressure      .68* 
Superheat              60° 
B.P.M.                 2500 

'n 

^ 

^ 

\ 

> 

*• 

\ 

** 

— 

^» 

•** 

^i 

^ 

\ 

^ 

^ 

^. 

^> 

\ 

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s 

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J 

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r, 

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50  GO  70 

Brake  Horse  Power 


30  40 

| 

FIG.  198a.     Typical  Performance  of  a  90  H.P.  Terry  Steam  Turbine. 


Any  number  of  individual  tests  may  be  cited  showing  superiority  in 
fuel  consumption  of  the  piston  engine  over  that  of  a  turbine  of  equiva- 
lent capacity  and  vice  versa,  but  when  the  machines  are  designed  for 


tbs.of  Steam. 
j?erK.W.Hour 

s  co  ^ 

=- 

—  •*. 

= 

—  -?• 

—  , 

— 

__i_ 

—  — 

=.  

5000    6000     7000     8000     9000    10000    11000    12000 
Load  in  K.W. 

FIG.  198b.     Typical  Performance  of  9000  Kw.  Curtis  Turbine;  200  Lbs.  Gauge 
Pressure,  125°  Superheat,  29  Inches  Vacuum. 

the  same  operating  conditions  the  results  are  practically  the  same  for 
all  sizes  under  2000  kw.  Tables  39  to  43  give  the  general  condition 
of  operation  and  the  steam  consumption  of  exceptionally  good  piston 
engines  of  various  sizes  and  types,  and  Table  48  similar  data  of  first- 


386 


STEAM  POWER  PLANT  ENGINEERING 


class  turbines.  A  study  of  these  tables  will  show  that  the  choice  must 
be  based  on  other  factors  than  the  steam  consumption.  In  a  general 
sense,  the  piston  engine  is  superior  to  the  turbine  for  high  back  pres- 
sures, slow  rotative  speeds  and  heavy  starting  torques,  while  the  tur- 
bine has  practically  superseded  the  engine  for  large  central  station 
units  and  for  auxiliaries  requiring  high  rotative  speed.  Recent  tests 
of  the  Melville  reduction  gear  (Machinery,  Feb.,  1910)  show  exception- 
ally high  efficiencies  for  sizes  as  large  as  6000  kilowatts,  and  it  is  not 


Steam  Press.- 150  Lb. 
Dry  Steam 
Atmospheric  Exhaust 


Load 
FIG.  198c. 

unlikely  that  the  turbine  equipped  with  this  device  will  offset  the  low 
rotative  speed  factor  of  the  piston  engine. 

If  the  tests  of  steam  turbines  and^piston  engine  could  be  made  at  some 
standard  initial  pressure,  back  pressure  and  quality  or  superheat,  then 
a  comparison  could  readily  be  made,  but  both  types  of  prime  movers 
are  designed  to  give  the  best  results  for  special  operating  conditions, 
and  any  marked  departure  from  these  conditions  will  result  in  loss  of 
economy.  It  is  frequently  desired,  however,  to  make  a  comparison 
between  the  economy  of  the  different  machines,  and  the  following 
methods  are  in  vogue: 

(1)  Steam  consumption  under  assumed  conditions. 

(2)  Heat  consumption  per  horse-power  per  minute  above  the  ideal 
feed-water  temperature. 

(3)  Efficiency  ratio  or  ratio  of  ideal  to  actual. 


STEAM   TURBINES 


387 


Standard  Correction  Curves. 

This  method  for  comparing  engines  or  turbines  or  both  is  best  illus- 
trated by  a  specific  example:  Suppose  it  is  required  to  compare  the 
full-load  performance  of  a  125-kw.  direct-connected  piston  engine 
with  that  of  a  125-kw.  turbo-generator  with  operating  conditions  as 
follows: 


Steam  Consump- 
tion, Lbs.  per 
Kw.-Hour. 

Initial  Pres- 
sure, Lbs. 
Absolute. 

Vacuum, 
Inches  of 
Hg. 

Superheat, 
Deg.  F. 

Engine  

25.0 

160 

25.5 

o 

Turbine 

22  7 

110 

28  0 

125 

Manufacturers  of  steam  turbines  have  provided  correction  curves  as 
illustrated  in  Fig.  194d,  showing  the  influence  of  varying  vacuum, 
superheat  and  pressures  on  the  steam  consumption.*  From  curve 
B,  we  find  that  the  steam  consumption  of  the  turbine  should  be 


100 


110 


120 


Steam  Pressure,  Lbs. per  Sq.  In.  -Absolute 
130  140  150  160  170 


180 


190 


200 


28 


'24 


TYPICAL  CORRECTION  CURVES 

FOR  125  K.W.   STEAM  TURBINE 

FULL  LOAD  CONDITIONS. 


A    OF  Superheat,165^Abs. 

B      »  »       ,   28  in  Vacuum 

C    165^  Absolute,     »    »         »» 


20  40  CO  80  100  120 

Superheat,  Deg.  Fan. 
20  21  *2  23  24  25 

Vacuum,  Inches  of  Mercury 

FIG.  198d. 


140 
26 


160 
27 


180 


200 
29 


decreased  2.5  pounds  to  give  the  equivalent  at  160  pounds  initial 
pressure;  from  curve  A  it  should  be  increased  2.5  pounds  to  give  the 
equivalent  at  25.5  inches  of  vacuum,  and  from  curve  C  it  should  be 
increased  2.5  pounds  to  give  the  equivalent  at  0  degree  superheat. 
The  full-load  steam  consumption  for  the  turbine  under  the  engine  con- 
ditions is  therefore  22.7  -  2.5  +  2.5  -f-  2.5  =  25.2  pounds  per  kw.-hour. 
*  These  curves  are  drawn  to  a  much  larger  scale  than  the  reproduction  given  here. 


388  STEAM  POWER  PLANT  ENGINEERING 

The  ratio  method  is  also  used  in  this  connection,  thus:  The  full-load 
steam  consumption  at   160  pounds  pressure,  curve   B,   Fig.   194d,  is 

25 

multiplied  by  the  ratio  — —  to  give  the  equivalent  consumption  at  110 
z/.o 

pounds  (25  is  the  steam  consumption  at  160  pounds  and  27.5  the  con- 
sumption at   110  pounds).     Similarly  the  correction  ratio  to  change 

25  5 
the  consumption  at  28  inches  of  vacuum  to  25.5  is  -77=-,  and  to  correct 

25 
125°  F.  superheat  to  0°  F.  is  —-• 


SUMMARY. 

Pressure  correction  7——  =  0.91  =  —   9%. 
ZY.o 

27  5 

Vacuum  correction-^  =  1.10  =     10%. 
Zo 

25 
Superheat  correction  jrj—  =  1.11  =     11%. 


Net  correction  12%. 

Corrected  steam  consumption  =  22.7  +  0.12  X  22.7  =  25.4  pounds 
per  kw.-hr. 

The  ratio  method  is  generally  used  if  the  difference  between  the 
corrected  steam  consumption  and  that  of  the  correction  curves  for 
the  same  conditions  is  greater  than  5  per  cent  ("The  Steam  Turbine," 
Moyer,  p.  128). 

This  ratio  method  for  correcting  steam  consumption  at  full  load  may 
be  used  without  appreciable  error  for  half  to  one  and  one  half  load  and 
is  the  only  practical  method  for  quarter  load  (Engrg.  London,  March  2, 
1906). 

Heat  Consumption. 

The  heat  consumption  B.T.U.  per  unit  output  per  minute  above 
the  ideal  feed  water  temperature  may  be  expressed 

If  (JET,-  .ft) 


60 
For  the  case  cited  above 


See  equation  (75b). 


Engine,  ^_L  .  455  B.T.U. 

Turbine,  22.7(1264.2-70)  =  ^  ^ 


STEAM  TURBINES 


389 


Efficiency  Ratio. 

The  efficiency  ratio,  or  the  extent  to  which  the  theoretical  possibilities 
are  realized,  may  be  expressed 

2545 


Er       W  (H,  -  H2 

For  the  case  cited  above 

2545 


See  equation  (76b). 


Engine, 
Turbine, 


25(1194.1  -  915) 

2545 

22.7  (1264.2  -  915.3) 


=  0.366. 


-  0.322. 


In  the  assumed  case  the  turbine  is  the  more  economical  in  heat  con- 
sumption, but  the  engine  is  the  more  perfect  of  the  two  as  far  as  theo-, 
retical  possibilities  are  concerned. 


»  .i  .6  .7 

Load  ia  ttrnv;.  of  ".ted  electrical  load 


TIG.  199. 


202.  First  Cost.  —  Steam  turbines,  generally  speaking,  are  about  10 
per  cent  lower  in  first  cost  than  high-grade  compound  engines  of  equiv- 
alent power.  The  following  table  gives  an  idea  how  the  price  varies 
with  the  conditions  of  operation.  The  figures  are  approximate  only 
and  refer  to  the  cost  of  the  turbine  and  generator  exclusive  of  auxil- 
iaries. 


390 


STEAM  POWER  PLANT  ENGINEERING 


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392 


STEAM  POWER  PLANT  ENGINEERING 


APPROXIMATE    COST    OF    STEAM    TURBINES    AND    GENERATORS. 
In  Dollars  per  Kilowatt.  Rated  Capacity. 


Kilowatts. 

25 

50 

75 

100 

200 

300 

500 

1000 

2000 

4000 

6000 

Direct  current: 
Non-condensing  .... 
Condensing  

55 
60 

47 
51 

42 

46 

38 
43 

32 
36 

32 
36 

35 
35 

32 
30 

28 
27 

Alternating  current: 
25  cycles  

25 
25 

21 
21 

20 
20 

60  cycles  

203.  Cost  of  Operation.  —  Data  pertaining  to  the  cost  of  operating 
steam-turbine  and  reciprocating  engine  plants  and  combinations  of 
both  will  be  found  in  Chapter  XVII.  The  following  table,  contributed 
by  H.  G.  Stott,  Superintendent  Motive  Power  of  the  Interborough 
Rapid  Transit  Company,  New  York,  gives  an  excellent  comparison  of 
the  relative  maintenance  and  operating  costs  (March,  1910)  of  the 
three  types  of  steam  power  plants  as  applied  to  large  central  stations 
for  electric  street  railways. 

RELATIVE    COSTS    PER    KILOWATT-HOUR.      DISTRIBUTION  OF 
MAINTENANCE    AND    OPERATION. 


Reciprocating 
Steam 
Plant. 

Steam 
Turbine 
Plant. 

Reciprocating 
Engines  and  Low- 
pressure   Steam 
Turbines. 

Maintenance. 
1.   Engine  room,  mechanical  
2.   Boiler  or  producer  room 

2.59 
4  65 

0.51 
4  33 

1.55 
3  55 

3.   Coal  and  ash  handling  apparatus.  .  .  . 
4.   Electrical  apparatus  

0.58 
1  13 

0.54 
1  13 

0.44 
1  13 

Operation. 
5     Coal 

61  70 

55  53 

46  48 

6.  Water 

7  20 

0  65 

0  61 

7.   Engine  room  labor     .        .          

6  75 

1  36 

4  06 

8.   Boiler  or  producer  room  labor  
9.   Coal  and  ash  handling  labor  

7.20 

2.28 

6.74 
2.13 

5.50 
1  75 

10.   Ash  removal 

1  07 

0  95 

0  81 

11.   Electrical  labor  

2  54 

2  54 

2  54 

12.   Engine  room  lubrication  

•1  78 

0  35 

1  02 

13.   Engine  room  waste,  etc  

0  30 

0  30 

0  30 

14.   Boiler  room  lubrication   etc         .    . 

0  17 

0  17 

0  17 

Relative  operating  cost,  per  cent.  .  .  . 
Relative  investment,  per  cent  
Probable  average  cost  per  kw 

100.00 
100.00 
125  00 

77.23 
75.00 
93  75 

69.91 
80.00 
100  00 

Probable  fixed  charges.  . 

11% 

11% 

11% 

For  steam  turbine  plants  larger  than  60,000  kw.  the  cost  per  kilowatt 
may  be  reduced  to  $75.00. 


STEAM  TURBINES 


393 


204.  Influence  of  Superheat.  —  The  use  of  superheated  steam  jn- 
creases  the  economy  of  the  reciprocating  engine  about  1  per  cent  for 
every  10  to  20  degrees  of  superheat,  depending  upon  the  conditions  of 
operation,  the  gain  being  due  mainly  to  the  reduction  of  cylinder  con- 
densation. Cylinder  condensation  is  reduced  not  only  because  of  the 
excess  heat  available  for  the  evaporation  of  moisture  but  also  because 
superheated  steam  has  a  lower  conductivity  than  wet  steam,  and  less 
heat  is  given  up  to  the  cylinder  walls  for  the  same  difference  of  tem- 
perature. In  the  steam  turbine  this  difference  of  temperature  is  much 
smaller,  since  high-  and  low-pressure  steam  do  not  alternately  come  in 


RELATION  SUPERHEAT  TO  ECONOMY 
400  K-W.  Westinghousc-Parsons  Turbine 

^.j  28  inches  Vacuum 

Tested  by  Messrs.  Dean  &  Main  Eng'ra. 


o       20       40        eo       80      loo       mm       ieo      iso. 


FIG.  200. 

contact  with  the  same  surface  as  is  the  case  with  the  reciprocating 
engine,  and  the  time  of  contact  is  considerably  less,  due  to  the  com- 
paratively high  velocities.  With  a  well-lagged  casing,  therefore,  the 
condensation  due  to  this  cause  is  insignificant  compared  with  that  of 
the  reciprocating  engine,  and  the  beneficial  effect  of  superheat  is  much 
more  pronounced.  Friction  of  the  steam,  which  in  the  reciprocating 
engine  is  negligible,  and  which  may  be  a  source  of  considerable  loss  in 
the  turbine,  is  greatly  reduced  by  the  use  of  superheated  steam,  as  is 
also  the  "  windage  "  loss  due  to  the  rapid  revolution  of  the  wheels. 

The  problem  of  cylinder  lubrication  is  sometimes  a  difficult  one  in 
steam  engines  using  a  high  degree  of  superheat,  and  trouble  is  fre- 
quently experienced  due  to  the  unequal  expansion  of  the  metal.  In 


394 


STEAM   POWER   PLANT  ENGINEERING 


the  steam  turbine  the  latter  difficulty  is  not  so  pronounced  and  no 
internal  lubrication  is  necessary,  hence  higher  degrees  of  superheat 
are  permissible.  For  maximum  economy  the  steam  at  the  end  of 
expansion  should  be  free  from  moisture.  Assuming  purely  adiabatic 
expansion,  the  steam  in  expanding  from  165  pounds  to  1  pound  abso- 
lute would  have  to  be  superheated  about  500  degrees  F.,  giving  the 
steam  an  actual  temperature  of  800  degrees  F.  A  study  of  some  100 
tests  made  in  this  country  gives  about  250  degrees  superheat  as  a  maxi- 


36 
.34 
32 
SO 
28 
26 
3A 
22 
20 

RELATION-  YACUUM  TO  ECONOMY 
300  K.W.  Parsons  Turbine 
Hulton  Colliery 

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FIG.  201. 


30 


mum  and  100  degrees  to  150  degrees  F.  as  an  average.  In  Europe 
reciprocating  engines  are  operating  with  superheat  as  high  as  350 
degrees  F.  and  turbines  300  degrees  F.  The  additional  fixed  and 
operating  costs  of  superheating  must  be  considered  in  determining  the 
net  gain,  since  the  decrease  in  steam  consumption  does  not  represent 
the  actual  saving.  With  pressures  of  175  pounds  gauge  or  less,  and 
not  to  exceed  200  degrees  F.  superheat,  the  net  gain  has  in  most  cases 
proved  a  substantial  one.  With  higher  temperatures  and  pressures  the 
cost  of  maintaining  the  superheat  may  increase  more  rapidly  than  the 
saving  in  steam  consumption,  until  a  limit  is  reached  beyond  which  no 


STEAM  TURBINES 


395 


advantage  is  gained.  The  relation  between  superheat  and  steam  con- 
sumption for  a  400-kilowatt  Westinghouse-Parsons  turbine  is  illustrated 
in  Fig.  200.  Fig.  202  gives  a  similar  comparison  for  a  1500-kilowatt 
turbine.  (J.  R.  Bibbins,  Power,  January,  1905.) 

205.  Influence  of  High  Vacua.  —  The  possible  economy  of  the  recip- 
rocating engine  is  greatly  restricted  by  its  limited  range  of  expansion. 
Cylinders  cannot  be  profitably  designed  to  accommodate  the  rapid 
increase  in  the  volume  of  steam  when  expanded  to  very  low  pressures. 
For  example,  the  specific  volume  of  1  pound  of  steam  under  a  vacuum 
of  29  inches  (referred  to  a  30-inch  barometer)  is  about  650  cubic  feet, 
or  nearly  double  its  volume  under  a  vacuum  of  28  inches.  Usually 
the  exhaust  is  opened  at  a  pressure  of  6  or  8  pounds  absolute  and 
consequently  a  large  proportion  of  the  available  energy  is  lost.  The 
lower  vacuum  in  the  exhaust  pipe,  therefore,  serves  only  to  diminish 
the  back  pressure  and  does  not  affect  the  completeness  of  expansion. 
Even  if  it  were  practical  to  expand  to  1  pound  absolute,  the  increased 
condensation  in  the  reciprocating  engine  would  offset  any  gain  due  to 
expansion  unless  the  steam  were  highly  superheated.  A  study  of  a 
number  of  tests  of  reciprocating  engines  shows  a  slight  improvement 
due  to  increasing  the  vacuum  beyond  26  inches.  Tests  of  steam  tur- 
bines show  a  decrease  of  3  to  4  per  cent  in  steam  consumption  for  each 
inch  increase  of  vacuum  between  25  and  29  inches,  for  with  a  well- 
lagged  casing  cylinder  conden-  _ 
sation  is  practically  absent, 
since  the  high-  and  low-temper- 
ature steam  do  not  alternately 
come  in  contact  with  the  metal- 
lic surf  aces  as  is  the  case  with  the 
reciprocating  engine  (Figs.  201 
and  201a).  Fig.  230  shows  a 
relation  between  the  power  con- 
sumption of  the  auxiliaries  and 
the  total  output  of  the  station 
at  different  loads  for  a  Parsons 
steam  turbine  installation  and 
Fig.  231  shows  a  similar  relation 
for  the  2000-kilowatt  Curtis  tur- 
bine. The  power  consumption 
in  the  latter  case  is  higher  on  account  of  the  high  temperature  of  cool- 
ing water.  Table  55  gives  the  power  required  for  the  auxiliaries  in  a 
number  of  stations.  A  high  vacuum  may  be  limited  by  the  initial  tem- 
perature of  the  cooling  water.  The  difference  in  temperature  between 


LATION-VACUUM  TO  ECONOMI 
Westinghouse-Parsons  Turbine 
No.  Superheat 


23   .2    .4    .6     .8     26    .2    A     .«     .8      27    .2     A     .6    .8  3» 


396 


STEAM  POWER  PLANT  ENGINEERING 


16 
15.5 

15 

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U 
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1X3 
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2 

}v 

Effect  of  Vtu 
on  Steai 

1500  K.Y 

iuum  and  Superheat 
m  Consiunption 

r.,  Turbine  ,FulI  Load 

p.;  — 

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5           26            27           28 

40    60     80    100  120  UD 
Superheat  Deg.E. 


FIG.  202. 


inlet  and  discharge  should  be  greater  than  10  degrees,  since  otherwise 
the  amount  of  circulating  water  per  pound  of  steam  becomes  excessive 
and  increases  the  work  of  the  pumps.  For  example,  the  temperature 
of  steam  corresponding  to  a  vacuum  of  28  inches  or  1  pound  absolute 

is  102  degrees,  and  with  cooling 
water  at  75  degrees  F.,  and  the 
discharge  at  95  degrees  F.,  the 
theoretical  ratio  of  cooling  water 
to  steam  necessary  to  maintain 
this  vacuum  will  be  about  50 
and  the  actual  nearer  70.  From 
Table  50  it  will  be  seen  that 
a  28-inch  vacuum  referred  to  a 
30-inch  barometer  is  obtained 
with  an  average  ratio  of  50 
pounds  of  cooling  water  at  70 

degrees  F.  per  pound  of  steam.  The  cost  of  high-vacuum  apparatus  is 
not  proportional  to  the  vacuum,  but  increases  much  more  rapidly,  as 
shown  in  Fig.  232.  These  estimates  show  averages  and  not  specific  costs. 
Fig.  202  shows  the  effect  of  superheat  and  vacuum  on  the  economy 
of  a  1500-kilowatt  Westinghouse-Parsons  turbine.  Figs.  200  to  202  are 
taken  from  "  Steam  Power  Plants,"  by  J.  R.  Bibbins,  as  published  in 
Power,  January,  1905. 

Reciprocating  Engine  vs.  Turbine:  Power,  April,  1904,  p.  232,  May,  1904,  p.  298; 
Engr.  U.  S.,  Nov.  1,  1905,  p.  711;  Elec.  World,  April  2,  1904,  p.  651;  Eng.  Mag., 
Sept.,  1905,  p.  935;  Power,  Feb.,  1906,  p.  83;  Elec.  Age,  June,  1905,  p.  478;  Elec. 
Rev.,  Dec.  23,  1904. 

Steam  Turbine  Design:  Eng.  Rec.,  July  22,  1905,  p.  101;  St.  Ry.  Jour.,  Dec.  20, 
1902,  p.  988;  Engr.,  Lond.,  Jan.  8,  1904,  p.  34,  May  13,  1904,  p.  481,  Dec.  27,  1907; 
Electrician,  Lond.,  March  24,  1905;  Power,  Dec.,  1905;  Mech.  Eng.,  Feb.  7,  1908; 
Engineering,  Dec.  13, 1907. 

Theory  and  Design  of  Steam  Turbines:  Engr.  U.  S.,  Dec.  16,  1907,  p.  1126  (serial), 
March  15,  p.  201;  Revue  de  Mecanique,  Oct.  31,  1907;  Engr.,  Lond.,  Oct.  4,  1907; 
Eng.  Rev.,  May,  1904;  Mech.  Engr.,  Oct.  28,  1905;  Eng.  Rec.,  July  22,  1905,  p.  101, 
May  7,  1904,  p.  581. 

Modern  Steam  Turbine  Plants:  Power,  Dec.,  1906,  p.  717,  Dec.,  1907;  Engr.  U.S., 
Nov.  15,  1906,  p.  733,  March  15,  1907,  p.  304;  St.  Ry.  Jour.,  Oct.  19,  1907;  Elec. 
World,  July  22,  1905,  Feb.  15,  1908;  Eng.  Rec.,  March  4,  1905. 

Tests  of  Westinghouse-Parsons  Turbines:  Engr.  U.S.,  Dec.  1,  1904,  p.  802;  St.  Ry. 
Jour.,  Dec.  19,  1903,  p.  1063;  Power,  March,  1904,  p.  127,  April,  1904,  p.  239,  Aug., 
1905,  p.  466;  Elec.  World,  Sept.  6,  1902,  p.  360;  Eng.  Rec.,  July  29,  1905,  p.  134. 

Governing  Steam  Turbines:  Harvard  Engineering  Jour.,  1908. 

A  Recent  Comparison  of  Turbines  and  Engines:  Eng.  Rec.,  Feb.  19,  1910. 

Internal  Losses  of  a  Steam  Turbine:  Power  &  Engr.,  Aug.  24,  1909. 

The  Principles  of  Steam-turbine  Buckets:  Power,  Mar.  17,  1908. 


CHAPTER  XI. 

CONDENSERS. 

206.  General.  —  A  pound  of  dry  steam  at  atmospheric  pressure  (30 
inches  mercury)  occupies  a  volume  of  26.8  cubic  feet.  Suppose  these 
26.8  cubic  feet  of  steam  were  contained  in  a  closed  vessel,  and  that  the 
steam  was  subsequently  condensed  and  its  temperature  lowered  by 
suitable  means  to  say  110  degrees  F.  The  condensed  steam  would 
occupy  only  about  T7\><y  °f  its  original  volume,  and  the  pressure  would 
fall  to  2.6  inches  of  mercury,  the  latter  pressure  being  due  to  the  ten- 
sion of  the  aqueous  vapor  at  the  given  temperature.  That  is  to  say, 
the  best  vacuum  theoretically  attainable  under  the  given  conditions 
would  be  30  —  2.6  =  27.4  inches.  The  lower  the  temperature  to 
which  the  condensed  steam  is  reduced  the  more  nearly  perfect  will  be 
the  vacuum  attained. 

If  air  is  mixed  with  the  steam  the  vacuum  will  be  still  more  imperfect. 
Thus,  suppose  the  vessel  to  contain  one  pound  of  steam  and  one-tenth  of 
a  pound  of  air  under  atmospheric  pressure.  The  volume  of  the  closed 
vessel  in  this  case  must  be  26.8  +  1.69  =  28.49  cubic  feet. 

After  the  steam  has  been  condensed  and  its  temperature  reduced  to 
110  degrees  F.  the  remaining  pressure  will  be  due  to  the  aqueous  vapor 
tension  plus  the  pressure  due  to  the  air,  or  2.6  +  1.51  =  4.11  inches 
mercury,  and  the  maximum  vacuum  attainable  will  be  25.89  inches. 
In  practice  air  is  always  present  in  exhaust  steam.  A  condenser  is  a 
device  in  which  the  process  of  condensation  and  subsequent  removal 
of  the  air  and  condensed  steam  is  continuous,  the  degree  of  vacuum 
obtained  depending  upon  the  tightness  of  valves  and  joints,  the  quan- 
tity of  entrained  air,  and  the  temperature  to  which  the  condensed  steam 
is  reduced.* 

The  degree  of  vacuum  may  be  expressed  in  different  ways.  (1)  Excess 
of  the  atmospheric  pressure  over  the  observed  vacuum.  For  example, 
a  26-inch  vacuum  implies  that  the  pressure  of  the  atmosphere  is  26 
inches  of  mercury  above  the  pressure  in  the  condenser.  (2)  Per  cent 
of  vacuum,  by  which  is  meant  the  ratio  of  the  observed  vacuum  to 
the  atmospheric  pressure.  Thus  with  the  barometer  standing  at  30 
inches  a  vacuum  of  26  inches  may  be  expressed  as  100  X  M  =  86.6 
per  cent  vacuum.  This  method  of  expression  gives  an  idea  of  the 

*  See  "  The  Influence  of  Air  on  Vacuum  in  Surface  Condensers,"  Engng.,  April 
17,  1908;  Power,  Feb.  2,  1909,  p.  235. 

397 


398  STEAM  POWER  PLANT  ENGINEERING 

efficiency  of  the  condensing  system.  For  example,  the  degree  of 
vacuum  indicated  by  26  inches  would  be  93  per  cent  with  a  barometric 
pressure  of  28  inches  but  only  84  per  cent  when  the  barometer  reads  31 
inches.  (3)  Absolute  pressure.  Thus  a  26-inch  vacuum  referred  to  a 
30-inch  barometer  would  be  indicated  as  a  pressure  of  30  —  26  =  4  inches 
absolute,  or  1.9  pounds  per  square  inch. 

307.  Function  of  the  Condenser.  —  The  function  of  a  condenser  in 
connection  with  a  steam  engine  or  turbine  is  primarily  the  reduction  of 
back  pressure,  though  in  some  instances,  notably  in  marine  work,  the 
recovery  of  the  condensed  steam  may  be  of  equal-  importance.  The 
advantages  to  be  gained  by  decreasing  back  pressure  may  be  most 
readily  illustrated  by  the  following  example:  A  non-condensing  engine 
taking  steam  at  a  pressure  of  100  pounds  absolute  and  cutting  off  at 
one-quarter  stroke  will  have,  theoretically,  a  mean  effective  pressure 
on  the  piston  of  44.6  pounds  per  square  inch,  the  back  pressure  being 
14.7  pounds  per  square  inch  absolute.  If  the  engine  exhausts  into  a 
condenser  against  a  26-inch  vacuum  (1.7  pounds  absolute)  the  mean 
effective  pressure  will  be  increased  to  44.6  +  (14.7  —  1.7)  =  57.6 
pounds  per  square  inch,  resulting  in  a  gain  in  power  which  may  be 
expressed 

-^  "    -,:,        H-p-  =  i$r  (82) 

in  which 

H.P.  =  horse  power  gained. 

Pr  —  reduction  in  back  pressure,  pounds  per  square  inch. 
A  =  area  of  the  piston  in  square  inches. 
S  =  piston  speed  in  feet  per  minute. 

If  P  =  mean  effective  pressure  on  the  piston  when  running  non-con- 
densing, the  percentage  of  increase  of  power  may  be  expressed 

Percent  =  100-^f.  (83) 

In  the  above  example  the  percentage  of  power  gained  would  be 

100  -i2-  =  29.2  per  cent. 
44.6 

The  actual  gain  due  to  the  use  of  the  condenser  would  be  much 
less  than  this,  depending  upon  the  type  of  engine  and  conditions  of 
operation,  as  shown  in  the  results  of  engine  performances  outlined  in 
Chapter  X. 


CONDENSERS 


399 


TABLE   49. 

PRESSURE  OF  AQUEOUS  VAPOR  IN  INCHES  OF  MERCURY  FOR  EACH  DEGREE  F. 

(Marks  and  Davis.) 


0° 

1° 

2° 

3° 

4° 

5° 

6° 

7° 

8° 

9° 

30° 

.180 

.188 

.195 

.203 

.212 

.220 

.229 

.238 

40° 

248 

.257 

.268 

.278 

.289 

.300 

.312 

.324 

.336 

.349 

50° 

.362 

.376 

.390 

.405 

.420 

.436 

.452 

.468 

.486 

.503 

60°  
70°    

.522 
.739 

.541 
-  .764 

.560 
.790 

.580 

.817 

.601 
.845 

.622 
.873 

.644 
.903 

.667 
.964 

.690 
.996 

.714 
1.03 

80° 

1.03 

1.06 

1.10 

1.13 

1.17 

1.21 

1.25 

1.30 

1.33 

1.37 

90°   

1.42 

1.46 

1.51 

1.55 

1.60 

1.65 

1.71 

1.76 

1.81 

1.87 

100°  
110°  
120°  

1.93 
2.60 
3.44 

1.98 
2.66 
3.53 

2.04 
2.74 
3.63 

2.11 

2.82 
3.74 

2.17 
2.90 
3.84 

2.24 
2.99 
3.95 

2.30 
3.07 
4.06 

2.37 
3.16 
4.17 

2.44 
3.25 

4.28 

2.51 
3.34 
4.40 

130°  

4.52 

4.64 

4.76 

4.89 

5.02 

5.16 

5.29 

5.43 

5.58 

5.73 

140° 

5  88 

6.03 

6.18 

6.34 

6.51 

6.67 

6.84 

7.02 

7.20 

7.38 

With  steam  turbines  the  advantage  gained  by  reduction  of  back 
pressure  is  more  marked  than  with  the  reciprocating  engine,  though 
theoretically  the  same  for  the  same  range  of  expansion.  Initial  con- 
densation, leakage  past  valves,  and  other  sources  of  loss  prevent  a 
reciprocating  engine  from  benefiting  from  a  good  vacuum  to  the  same 
extent  as  a  turbine.  (See  paragraph  205.) 

Referring  again  to  the  example  given  above,  if  the  steam  is  cut  off  at 
about  one-sixth  stroke,  the  work  done  when  running  condensing  will  be 
the  same  as  when  running  non-condensing  and  cutting  off  at  one-quarter. 
Theoretically  the  steam  consumption  will  be  decreased  nearly  in  pro- 
portion to  the  reduction  in  cut-off.  Generally  speaking,  a  condensing 
engine  will  require  from  20  to  30  per  cent  less  steam  for  a  given  power 
than  a  non-condensing  engine.  (See  results  of  engine  tests,  paragraph 
179.)  This  decrease  in  steam  consumption  is  only  an  apparent  one.  If 
steam  is  used  by  the  auxiliaries  in  creating  the  vacuum,  the  amount  must 
be  added  to  that  consumed  by  the  engine,  unless  the  steam  exhausted 
by  the  former  is  utilized  to  warm  the  feed  water,  in  which  case  only  the 
difference  between  the  heat  entering  the  auxiliaries  and  that  returned 
to  the  heater  should  be  charged  against  the  engine.  The  power  neces- 
sary to  operate  the  condenser  auxiliaries  varies  from  one  to  six  per  cent 
of  the  main  engine  power,  depending  upon  the  type  and  conditions  of 
operation.  (See  paragraph  228.) 

In  power  plants  where  the  exhaust  steam  is  not  used  for  heating  or 
manufacturing  purposes,  the  engines  are  almost  invariably  operated 
condensing,  provided  there  is  an  abundant  supply  of  cooling  water. 
Even  if  the  water  supply  is  limited,  it  is  often  found  to  be  economical  to 


400 


STEAM  POWER  PLANT  ENGINEERING 


use  some  artificial  cooling  device,  notwithstanding  the  high  first  cost 
and  cost  of  operation  of  the  latter. 

Some  of  the  considerations  affecting  the  propriety  of  running  con- 
densing and  the  choice  of  condensing  systems  are  taken  up  in  para- 
graphs 230  and  231. 

The  Law  of  Condensation  of  Steam  :  Pro.  Inst.  of  Civil  Engrs.,  Nov.  30,  1897; 
Engr.,  Lond.,  Dec.  17,  1897,  p.  609.  Relation  of  Pressure  and  Temperature  in  Con- 
densers :  Power,  April,  1902,  p.  28,  June,  1902,  p.  30.  The  Measurement  of  Vacuum  : 
Engr.,  Lond.,  April  21,  1905.  Experiments  on  Condensation  of  Steam  :  Engr.,  Oct. 
15,  1897,  p.  481.  Condensation  Fallacies  and  Facts:  Machinery,  Sept.,  1904,  p.  38; 
Elec.  Review,  June  17,  1904.  Condenser  Pressure,  a  Neglected  Point  in  Steam 
Condensers :  Mech.  Engr.,  Sept.  30,  p.  484. 

Advantages  of  Condensing  :  Amer.  Elecn.,  Sept.,  1904,  p.  469.  Condensers  for 
Steam  Engines  and  Turbines  (F.  Foster) :  Mech.  Engr.,  Oct.  28,  1905,  pp.  637,  655. 
Condensers,  to  What  Extent  Should  They  be  Used :  Power,  Nov.,  1899.  The  Value 
of  the  Condenser :  Power,  Oct.,  1902.  The  Effects  of  Vacuum  on  Steam  Engine 
Economy:  Eng.  Mag.,  June,  1905.  Importance  of  Condensers:  Power,  May,  1897; 
Evaporating,  Condensing,  and  Cooling  Apparatus  (E.  Hausbrand),  published  by 
Scott,  Greenwood  &  Co.,  London. 

The  Use  of  Condensers  :  Electrician,  Lond.,  Oct.  12,  1900. 


1.  Jet  condensers. 


2.  Surface  condensers. 


Parallel  current  (a)... 


Counter  current  (6).. 


Water  cooled  (a) ... 


Ordinary  (1) 
Siphon  (2)... 


208.    Classification  of  Condensers.  —  The  following  is  a  classification 
of  a  few  well-known  steam  engine  condensers  : 

Worthington. 

Blake. 

Deane. 

Baragwanath. 

Bulkley. 

Ejector  (3) Schutte. 

f  Weiss. 

Barometric -I  Alberger. 

[  Tomlinson. 

Single-flow Baragwanath. 

Double-flow Wheeler. 

Multi-flow  Wainwright. 

Forced  draft Fouche. 

Natural  draft Pennell. 


Air  cooled  (6) 


Evaporative  (c) .Ledward. 


Condensers  may  be  divided  into  two  general  groups: 

1.  Jet  condensers,  in  which  the  steam  and  cooling  water  mingle  and 
the  steam  is  condensed  by  direct  contact,  Figs.  203  to  211. 

2.  Surface  condensers,  in  which  the  steam  and  cooling  medium  are 
in  separate  chambers  and  the  heat  is  abstracted  from  the  steam  by  con- 
duction, Figs.  212  to  215. 


CONDENSERS  401 

Jet  condensers  may  be  further  grouped  into  two  classes,  according  to 
the  direction  of  flow  of  the  air  and  cooling  water: 

(a)  Parallel-current  condensers,  in  which  the  condensed  steam,  cool- 
ing water,  and  air  flow  in  the  same  direction,  collect  at  the  bottom  of 
the    condenser  chamber,    and    are    exhausted  by   a   suitable    pump, 
Fig.  203. 

(b)  Counter-current  condensers,  in  which  the  cooling  water  and  con- 
densed steam  flow  from  the  bottom  of  the  chamber,  usually  by  gravity, 
while  the  air  is  drawn  off  at  the  top,  Fig.  229b. 

Parallel-current  condensers  may  be  subdivided  into  three  classes: 

(1)  Standard  condensers,   in  which   the   cooling    water,   condensed 
steam,   and  air   are  exhausted  by   a   vacuum  pump,   Figs.   203  and 
209. 

(2)  Siphon  condensers,  in  which  the  cooling  water,  condensed  steam, 
and  air  are  exhausted  by  a  barometric  column,  Fig.  205. 

(3)  Ejector    condensers,    in    which    the    condensed    steam    and    air 
are  exhausted  by  the  cooling  water,  on  the  ejector   principle,  Fig. 
206. 

Surface  condensers  may  be  classified  according  to  the  nature  of  the 
cooling  medium  as 

(a)  Water-cooled  condensers,  Fig.  212. 

(b)  Air-cooled  condensers,  Fig.  217.. 

(c)  Evaporative  condensers,  in  which  the  condensation  of  the  steam  is 
brought  about  by  the  evaporation  of  a  fine  stream  of  water  trickling  on 
the  outside  of  the  tubes. 

209.  Ordinary  Jet  Condensers.  —  Fig.  203  shows  a  section  through  a 
Worthington  jet  condenser,  illustrating  the  parallel-current  principle. 
When  the  pump  is  started  a  partial  vacuum  is  created  in  the  suction 
chamber  above  the  valves  H,  H  in  the  cone  F.  As  soon  as  sufficient 
air  has  been  exhausted,  cooling  water  enters  at  B  with  a  velocity  depend- 
ing upon  the  degree  of  vacuum  in  chamber  F  and  the  suction  head, 
and  is  divided  into  a  fine  spray  by  the  adjustable  serrated  cone  D. 
The  spray  mingles  with  the  exhaust  steam  entering  at  A  and  both 
move  downwards  with  diverse  velocities.  The  steam  gives  up  its  heat 
to  the  water  and  condenses.  The  velocity  of  the  steam  diminishes  in  its 
downward  path  to  zero,  while  the  velocity  of  the  water  increases  accord- 
ing to  the  laws  of  falling  bodies.  The  condensed  steam,  cooling  water, 
and  air  collect  at  the  lower  part  of  the  condenser  and  are  exhausted  by 


402 


STEAM   POWER  PLANT  ENGINEERING 


the  wet  air  pump  G,  from  which  they  are  forced  through  opening  J  to 
the  hot  well.  The  vacuum  in  chamber  F  will  depend  upon  the  vapor 
tension  of  the  warm  water  in  the  bottom  of  the  well,  the  amount  of 
air  carried  along  by  the  cooling  water  and  steam,  and  the  tightness  of 


FIG.  203.     Worthington  Independent  Jet  Condenser. 

valves  and  joints.  In  case  the  water  accumulates  in  the  condenser 
cone  F,  either  by  reason  of  an  increased  supply  or  by  a  sluggishness  or 
even  stoppage  of  the  pump,  the  condensing  surface  is  reduced  to  a  mini- 
mum, as  soon  as  the  level  of  the  water  reaches  the  spray  pipe  and  the 
spray  becomes  submerged,  and  only  a  small  annular  surface  of  water  is 
exposed  to  the  exhaust  steam.  The  vacuum  is  immediately  broken, 
and  the  exhaust  steam  escapes  by  blowing  through  the  injection  pipe 


CONDENSERS 


403 


and  through  the  valves  of  the  pump  and  out  the  discharge  pipe  at  J, 
forcing  the  water  ahead  of  it;  consequently  flooding  of  the  steam 
cylinder  cannot  occur.  In  starting  up  the  condenser  a  partial  vacuum 
for  inducing  a  flow  of  injection  water  into  the  condenser  chamber  may 
be  created  by  the  pump  if  the  suction  lift  is  not  too  great.  Many 
engineers,  however,  prefer  to  install  a  small  forced  injection  or  priming 
pipe  the  function  of  which  is  to  condense  sufficient  steam  to  produce 
the  necessary  partial  vacuum.  Fig.  222  shows  such  an  installation. 


FIG.  204.     Section  through  a  Blake  Jet  Condenser. 

Fig.  204  shows  a  section  through  the  condensing  chamber  and  air 
pump  of  a  Blake  vertical  jet  condenser  with  an  automatic  vacuum- 
breaking  device.  The  injection  water  enters  at  opening  marked 
"  injection  "  and  flows  through  the  adjustable  "  spray  "  nozzle  in  a  fine 
spray,  at  an  angle  of  about  45  degrees,  and  impinges  on  the  conical  sides 
of  the  upper  condenser  chamber.  The  spray  falls  from  the  sides  to  the 
projecting  ledges  shown  in  the  illustration.  The  ledges  prevent  the 


404  STEAM  POWER  PLANT  ENGINEERING 

spray  from  falling  directly  to  the  bottom  of  the  chamber  and  insure  an 
efficient  mingling  of  steam  and  cooling  water.  A  perforated  copper 
plate  is  substituted  for  the  shelves  when  the  force  of  the  injection  water 
is  not  sufficient  to  produce  spray.  The  circulating  water  and  con- 
densed steam  together  with  the  non-condensable  gases  are  drawn  off  at 
the  bottom  of  the  chamber.  The  vacuum-breaking  device  is  shown 
at  the  right  of  the  figure.  When  the  rising  water  reaches  the  level  of 
the  float  chamber,  as  in  the  case  of  an  accidental  stoppage  of  the  air 
pumps,  the  float  is  raised  and  forces  a  check  valve  from  its  seat  and 
allows  an  inrush  of  air  to  break  the  vacuum,  thus  preventing  further 
suction  of  water  into  the  condenser  and  consequent  flooding  of  the 
engine.  A  is  the  forced  injection  or  "priming"  inlet  used  in  start- 
ing up  when  the  suction  lift  is  considerable. 

Condenser  Types  and  Applications:  Power,  June,  1906,  p.  44;  Engr.  U.S.,  Jan., 
1906,  pp.  55-66.  Jet  Condensers  (McBride):  Trans.  A.S.M.E.,  Vol.  12,  p.  187; 
American  Machinist,  March  7,  1895,  p.  185;  Engr.  U.S.,  Jan.,  1906,  p.  61;  Whitham's 
Steam  Engine  Design,  p.  294;  Seaton's  Manual  of  the  Marine  Engine,  Chapter  XI. 

210.  Condensing  Water,  Jet  Condensers.  —  In  a  jet  condenser  the 
cooling  water  and  exhaust  steam  mingle,  and  the  degree  of  vacuum  is 
a  function  of  the  final  or  discharge  temperature;  thus  the  quantity  of 
cooling  water  required  depends  upon  its  initial  temperature,  the  tem- 
perature of  the  discharge  water,  and  the  total  heat  in  the  steam  entering 
the  condenser.  If  the  steam  in  the  low-pressure  cylinder  at  release  is 
dry  and  saturated,  the  heat  entering  the  condenser  will  correspond 
to  the  total  heat  in  steam  at  release  pressure,  but  it  usually  contains 
considerable  moisture,  part  of  which  is  reevaporated  when  the  exhaust 
valve  opens  to  the  condenser;  however,  it  is  sufficiently  accurate  for 
all  practical  purposes  to  assume  the  exhaust  steam  entering  the  con- 
denser to  be  dry  and  saturated  and  its  heat  to  correspond  to  the  pres- 
sure in  the  condenser. 

Let  A.  =  total  heat  of  steam  at  condenser  pressure  above  32  degrees. 
T2  =  temperature  of  the  discharge  water. 
T0  =  initial  temperature  of  the  cooling  water. 
W  =  weight  of  cooling  water  in  pounds  necessary  to  condense 
and  cool  one  pound  of  steam  to  the  required  discharge 
temperature. 

Then  TF^-^t,32*  (84) 

•*2~"    -*0 

Example:  How  many  pounds  of  cooling  water  are  necessary  to  con- 
dense one  pound  of  steam  under  the  following  conditions:  Barometer 
29.92;  vacuum  26  inches;  temperature  of  injection  water  60  degrees  F. 


CONDENSERS  405 

The  temperature  of  aqueous  vapor  corresponding  to  an  absolute 
pressure  of  29.92  —  26  =  3.92  inches  of  mercury  is  125  degrees  F.  (See 
Table  49.)  The  discharge  temperature,  however,  must  be  less  than  this, 
as  the  pressure  in  the  condenser  is  due  not  only  to  the  aqueous  vapor  but 
to  that  of  the  air  carried  over  with  the  circulating  water  and  the  con- 
densed steam.  In  a  condenser  of  this  type  the  discharge  temperature 
will  be  from  10  degrees  to  15  degrees  lower  than  that  corresponding  to 
the  vacuum  as  recorded  by  the  gauge.  In  this  case  assume  it  to  be  15 
degrees  lower,  i.e.,  T2=  125  —  15  =  110  degrees. 

The  total  heat  corresponding  to  a  pressure  of  3.92  inches  of  mercury 
is  1120  B.T.TJ.  above  32  degrees  (see  steam  tables);  T7^  60  degrees; 
TZ  =  110  degrees. 


110-60 

Evidently  the  higher  the  temperature  of  the  discharge  water  the  less 
will  be  the  quantity  of  cooling  water  required,  and  consequently  the 
smaller  the  weight  of  air  introduced  into  the  condenser;  but  the  warmer 
the  discharge  water  the  greater  will  be  the  vapor  tension  and  the  lower 
the  degree  of  vacuum.  For  reciprocating  engines  a  hot-well  tempera- 
ture between  110  degrees  and  130  degrees  F.  is  average  practice;  with 
turbines  the  temperature  ranges  between  80  degrees  and  100  degrees  F. 
On  account  of  the  inefficient  heat  absorption  in  practical  installations, 
from  5  per  cent  to  15  per  cent  is  added  to  the  theoretical  weight  of  cool- 
ing water  as  determined  from  equation  (84).  Table  50  has  been  calcu- 
lated from  equation  (84). 

Cooling  Water  for  Condensers:  Am.  Mach.,  May  18,  1905,  p.  656;  Evaporation 
and  Condensing  Apparatus,  Hausbrand,  pp.  227,  240,  301,  318;  Steam  Power  Plants, 
Meyer,  p.  106. 

Wet-  Air  Pump,  Jet  Condensers.     (See  paragraph  285.) 

Circulating  Pumps.     (See  paragraph  297.) 

211.   Effect  of  Aqueous  Vapor  upon  the  Degree  of  Vacuum.  —  The 

futility  of  attempting  to  better  the  vacuum  by  exhausting  the  vapor  is 
best  illustrated  by  a  specific  problem. 

Required  the  volume  of  aqueous  vapor  to  be  withdrawn  per  hour 
from  a  condenser  operating  under  the  following  conditions,  in  order 
that  the  vacuum  may  be  increased  one  pound  per  square  inch:  Tem- 
perature of  discharge  water  125  degrees;  corresponding  vapor  tension 
4  inches  of  mercury;  barometer  30  inches;  relative  vacuum  26  inches; 
horse  power,  100;  steam  consumption  20  pounds  per  horse-power  hour; 
cooling  water  25  pounds  per  pound  of  steam  condensed. 

100  X  20  X  25  =  50,000  pounds  of  cooling  water  per  hour. 
=  833  pounds  of  cooling  water  per  minute. 


TABLE   50. 

RATIO,  BY  WEIGHT,  OF  COOLING  WATER  TO  STEAM  CONDENSED  (THEORETICAL). 

(Barometer  29.92.) 


Vacuum  24". 

Vacuum  25". 

Temperature  of  Steam  141°. 

Temperature  of  Steam  134°. 

of  In- 
jection. 

Temperature  of  Hot  Well. 

of  In- 
jection. 

Temperature  of  Hot  Well. 

HO 

115 

120 

125 

130 

105 

110 

115 

120 

125 

40 

15.0 

13.9 

12.9 

12.1 

11.4 

40 

16.1 

14.9 

13.8 

12.9 

12.1 

50 

17.5 

16.0 

14.8 

13.7 

12.8 

50 

19.0 

17.4 

16.0 

14.8 

13.7 

60 

21.0 

18.9 

17.3 

15.8 

14.6 

60 

23.2 

20.9 

18.9 

17.2 

15.8 

70 

26.2 

23.2 

20.7 

18.7 

17.1 

70 

30.0 

26.1 

23.0 

20.7 

18.7 

80 

35.0 

29.8 

25.9 

23.0 

20.5 

80 

42.0 

34.8 

29.6 

25.9 

22  8 

90 

52.4 

49.7 

34.6 

29.5 

25.6 

90 

70.0 

52.1 

41.5 

34.5 

29.4 

Vacuum  26//. 

Vacuum  21"  . 

Temperature  of  Steam  125°. 

Temperature  of  Steam  114°. 

Temp, 
of  In- 
jection. 

Temperature  of  Hot  Well. 

of  In- 
jection. 

Temperature  of  Hot  Well. 

100 

105 

110 

115 

90 

95 

100 

105 

40 

17.5 

16.1 

14.8 

13.8 

40 

21.2 

19.1 

17.4 

16.0 

50 

21.0 

19.0 

17.4 

16.0 

50 

26.5 

23.4 

20.9 

19.0 

60 

26  3 

23  2 

20  9 

18  8 

60 

35  3 

30  1 

26  2 

23  2 

70 

35.0 

30.0 

26.0 

23.0 

70 

52.9 

42.1 

34.9 

29.8 

80 

57.6 

42.0 

34.7 

29.6 

80 

52.3 

41.5 

Temp, 
of  In- 
jection. 

Vacuum  27.5". 
Temperature  of  Steam  108°. 

Temp, 
of  In- 
jection. 

Vacuum  28". 
Temperature  of  Steam  100°. 

Temperature  of  Hot  Well. 

Temperature  of  Hot  Well. 

80 

85 

90 

95 

75 

80 

85 

90 

40 
50 
60 
70 

26.6 
35.6 
52.3 

23.6 
30.3 
42.5 
70.8 

21.1 
26.4 
35.2 
52.8 

19.1 
23.4 
30.0 
42.0 

40 
50 
60 
70 

30.5 

42.7 
71.2 

26.6 
35.5 
53.2 

23.5 
30.2 
42.3 
70.6 

21.1 
26.3 
35.1 
52.7 



Temp, 
of  In- 
jection. 

Vacuum  28.5". 
Temperature  of  Steam  90°. 

Temp, 
of  In- 
jection. 

Vacuum  29". 
Temperature  of  Steam  77°. 

Temperature  of  Hot  Well. 

Temperature  of  Hot  Well. 

60 

65 

70 

75 

55 

60 

65 

67 

35 
40 
45 
50 

42.2 
54.0 
72.0 

35.8 
43.0 
53.5 
72.0 

30.6 
35.6 
42.8 
53.5 

29.2 
33.4 
38.8 
46.6 

35 
40 
45 
50 

52.0 
69.3 

43.0 
54.0 
71.5 

35.8 
43.0 
54.0 
72.0 

33.4 

38.4 
47.0 
61.0 



CONDENSERS  407 

Now  to  increase  the  vacuum  one  pound  per  square  inch,  approxi- 
mately 2  inches  of  mercury,  the  temperature  of  the  water  must  be 
lowered  to  102  degrees  F.,  that  is,  833  (125-102)  =  19,159  B.T.U. 

19  159 

must  be  abstracted  from  the  water  in  one  minute,  or       '        =  18.6 

1030 

pounds  of  water  to  be  evaporated  per  minute.  (1030  =  average  heat 
of  vaporization  of  water  under  26  to  28  inches  of  vacuum.)  Now,  one 
pound  of  vapor  at  102  to  125  degrees  F.  has  an  average  volume  of  270 
cubic  feet. 

Therefore  18.6  X  270  =  5022  cubic  feet  of  vapor  must  be  exhausted 
per  minute  to  increase  the  vacuum  from  26  to  28  inches,  which  is  man- 
ifestly impracticable. 

212.  Injection  Orifice.  —  The  velocity  of  water  entering  a  jet  con- 
denser, neglecting  friction,  may  be  determined  from  the  formula 

V  =  Vtgti,  (85) 

where 

V  =  velocity  of  the  water  in  feet  per  second. 
g  =  acceleration  of  gravity  =  32.2. 
h  =  total  head  in  feet. 

If       p  =  pressure  below  the  atmosphere  in  pounds  per  square  inch, 
/&!=  distance  in  feet  between   the   source  of   supply   and  the 
injection  orifice, 

then  h  =  2.3  p  ±  hv  (86) 

and  equation  (85)  may  be  written 


V=  8.025  \/2.3p±  ht.  (87) 

If  the  supply  is  under  pressure,  hl  is  positive;  if  under  suction,  it  is 
negative. 

Example:  What  is  the  theoretical  velocity  of  water  entering  a  con- 
denser with  26-inch  vacuum  (referred  to  30-inch  barometer);  suction 
head  8  feet 

Here   p  =  pressure    in   pounds    per   square   inch,  corresponding   to 
26  inches  of  mercury  =  12.8  pounds  per  square  inch. 

fc1=  8.  

V  =  8.025  V2.3  X  12.8-8 
=  37.1  feet  per  second 
=  2226  feet  per  minute. 

In  proportioning  the  injection  orifice  in  practice  the  maximum 
velocity  of  flow  is  assumed  to  be  between  1500  and  1800  feet  per  minute, 


408 


STEAM  POWER  PLANT  ENGINEERING 


or,  approximately,  area  of  injection  orifice  in  square  inches  =  weight  of 
injection  water  in  pounds  -T-  650  to  780.  ("  Manual  of  Marine  Engineer- 
ing/' Seaton,  p.  204.)  A  rough  rule  gives  area  of  orifice  =  area  of  low 
pressure  piston  in  square  inches  -r-  250.  (Seaton,  p.  204.) 

313.  Volume  of  the  Condenser  Chamber.  —  According  to  Thurs- 
ton  the  volume  of  a  jet  condenser  should  be  from  one-fourth  to  one- 
half  that  of  the  low-pressure  engine  cylinder.  ( "  Steam  Engine 
Manual,"  Thurston,  II,  127.) 

According  to  Hutton  the  volume  should  not  be  less  than  that  of  the  air 
pump  and  should  approximate  three- 
fourths  of  that  of  the  engine  cylinder 
in  communication  with  it. 

214.  Injection  and  Discharge  Pipes. 
— In  practice  the  diameter  of  the  injec- 
tion pipe  is  based  on  a  velocity  of  400 
to  600  feet  per  minute  and  that  of  the 
discharge  pipe  of  200  to  400  feet  per 
minute;    the  lower  figures    for  pipes 
under  8  inches  in  diameter,  the  upper 
range  for  larger  diameters. 

(Atmospheric     relief    valves.  —  See 
paragraph  351.) 

215.  Siphon  Condensers.  —  Fig.  205 
shows  a  section  through  a  Baragwa- 
nath    siphon    condenser,    illustrating 
the  principles  of  a  parallel-current  baro- 
metric condenser.     The  cooling  water 
enters  the  side  of  the  condenser  cham- 
ber at  A  and  passes  downward  in  a 
thin  annular  sheet  around  the  hollow 
cone  D.     The  exhaust  steam  enters  at 
B  and  is  given  a  downward  direction 
by  the  goose  neck  C.     It  flows  through 

the  nozzle  D  and  is  condensed  within  the  hollow  cone  of  moving  water, 
the  combined  mass  including  the  entrained  air  discharging  through  the 
contracted  throat  E  at  high  velocity  into  the  tail  pipe  F.  The  water 
column  in  the  tail  pipe  must  be  enough  to  overcome  the  pressure  of  the 
atmosphere;  i.e.,  it  should  be  34  feet  or  more  above  the  surface  of  the 
hot  well,  otherwise  water  would  rise  within  this  pipe  to  a  height  corre- 
sponding to  that  of  the  barometer,  which  is  approximately  34  feet  for  a 
barometric  pressure  of  30  inches  of  mercury.  This  is  not  strictly  true 
when  the  condenser  is  in  full  operation,  as  the  injector  effect  of  the 


FIG.  205.     Baragwanath  Siphon 
Condenser. 


CONDENSERS 


409 


moving  mass  is  sufficient  to  overcome  several  pounds  pressure,  and  the 
tail  pipe  may  be  less  than  34  feet,  but  to  provide  against  any  possibility 
of  the  water  being  drawn  into  the  cylinder  of  the  engine  the  length  is 
made  greater  than  34  feet.  The  spray  cone  D  is  adjustable  and  admits 
of  close  regulation  of  the  water  supply  without  changing  the  annular 
form  of  the  stream.  The  condensing  water  may  be  supplied  under 
pressure  or  under  suction.  For  lifts  not  greater  than  15  feet  no  supply 
pump  is  necessary,  the  water  being  raised  by  the  siphon  action  of  the 
condenser.  This  condenser  requires  the  same  amount  of  cooling  water 
per  pound  of  steam  as  the  standard  jet  condenser,  and  is  capable  of 
maintaining  a  vacuum  of  from  24  to  25  inches.  A  vacuum  of  28 ^  inches 
has  been  recorded  for  a  condenser  of  this  general  type.  (Trans. 
A.S.M.E.,  26-388.)  An  atmospheric  relief  valve  G  is  provided  in 
case  the  vacuum  fails  from  any  cause,  which  will  permit  the  steam  to 
escape  to  the  atmosphere. 

The  above  type  of  condenser  is  adapted  to  very  muddy  cooling  water, 
since  no  nitration  is  necessary  beyond  the  removal  of  such  solid  matter 
as  may  clog  up  the  annular  space  H. 

In  the  Armour  Glue  Works  at  Chicago  condensers  of  this  type  are 
successfully  maintaining  a  90  per  cent  vacuum  with  cooling  water  at 
60  degrees  F.,  and  the  circulating  water  is  practically  liquid  mud. 

Siphon  Condensers,  Discussion:  Trans.  A.S.M.E.,  Vol.  26,  p.  388.  Siphon  Con- 
densers: Electrical  World,  June,  1897,  p.  818;  Hutton,  The  Mechanical  Engineering 
of  Power  Plants,  p.  106;  Engr.  U.S.,  Jan.,  1906. 

216.  Size  of  Siphon  Condensers.  —  The  size  of  siphon  is  indicated  by 
the  diameter  of  the  engine  exhaust  pipe. 

Table  51  gives  the  sizes  of  barometric  condensers  as  manufactured 
by  prominent  makers. 

TABLE  51. 
SIZE    OF    SIPHON    CONDENSERS. 


Steam  to  be  Condensed. 

Steam  to  be  Condensed. 

Size  Usually 

Size  Usually 

Pounds  per 
Hour. 

Pounds  per 
Minute. 

Furnished, 
Inches. 

Pounds  per  Hour. 

Pounds  per 
Minute. 

Furnished, 
Inches. 

2,000 

33 

5 

8,000 

133 

10 

3,000 

50 

7 

10,000 

166 

12 

4,000 

66 

8 

15,000 

250 

14 

5,000 

83 

9 

20,000 

333 

14 

6,000 

100 

9 

Vacuum  26  inches;  barometer  30  inches. 


410 


STEAM  POWER  PLANT  ENGINEERING 


The  diameter  of  the  throat   may  be  closely  approximated  by  the 
empirical  formula 

Diam.  in  inches  =  0.0077  VWw,  (88) 

in  which 

W  =  weight  of  steam  to  be  condensed  per  hour. 

w    =  weight  of  water  required  to  condense  one  pound  of  steam. 

The  maximum  width  of  the  annular  opening  for  the  admission  of 
water  may  be  obtained  from  the  empirical  formula 

Ww 


Width  in  inches  = 


(89) 


STEAM 


EXHAUST 


DISCHARGE 


FIG.  206.     Schutte  Ejector 
Condenser. 


39,550  d 
in  which 

d  =  diameter  of  the  nozzle  or  bottom  of 

the  cone  in  inches. 
W  and  w  as  in  (88) 

217.  Ejector  Condenser.  —  Fig.  206  shows 
a  section  through  a  Schutte  exhaust  steam 
"induction"  condenser,  illustrating  the  prin- 
ciples of  the  ejector  condenser  in  which  the 


FIG.  207.     Piping  for  Schutte  Ejector 
Condenser. 


momentum  of  flowing  water  ejects  the  discharge  without  the  aid  of  the 
circulating  pump.  Exhaust  steam  enters  the  ejector  through  the  open- 
ing marked  "  exhaust,"  passes  through  a  series  of  inclined  orifices  and 


CONDENSERS 


411 


nozzles  at  considerable  velocity,  and,  meeting  the  cooling  water  in  the 
inner  annular  chamber,  is  condensed.  The  cooling  water  is  drawn  in 
continuously  through  the  opening  marked  "  water,"  by  virtue  of  the 
vacuum  formed,  and  sufficient  velocity  is  imparted  to  the  jet  to  dis- 
charge the  combined  mass  of  condensed  steam,  cooling  water,  and  air 
against  the  pressure  of  the  atmosphere. 

Adjustment  for  capacity  is  effected  by  raising  or  lowering  the  ram  R 
by  means  of  the  wheel  H.  An  adjustable  sleeve  controls  the  avail- 
able area  of  the  exhaust  inlet  by  covering  more  or  less  openings  in  the 
combining  tube.  When  the  cooling  water  is  supplied  under  pressure 
the  openings  marked  "  steam  "  and  0  are  blanked.  When  water  is 
taken  under  suction  and  water  under  pressure  is  available  for  starting, 
0  is  blanked  and  opening  marked 
" steam"  is  connected  with  the  pres- 
sure supply.  When  water  is  taken 
under  high  suction  and  live  steam  is 
used  for  starting,  inlet  marked  "  steam" 
is  connected  to  live  steam  and  an  over- 
flow check  valve  is  placed  at  0.  Fig. 

207  gives  an  outline  of  the  necessary 
piping  for  a  condenser  installation  of 
this  type.    These  condensers  are  made 
in  all  sizes   conforming  with  exhaust 
pipe  diameters    of    1J    to    20   inches. 
The  same  amount  of  cooling  water  is 
required    as   for   jet   condensing   and 
vacua  of  20  to  25  inches  are  readily 
obtained. 

Exhaust  Steam  Induction  Condensers  : 
Power,  Dec.,  1898,  p.  14.  Ejector  Condenser: 
Hutton,  Mechanical  Engineering  of  Power 
Plants,  p.  Ill;  Eng.  News,  Oct.  5,  1905, 
p.  360. 

218.   Barometric  Condensers.* — Fig. 

208  shows  a  section  through  a  Weiss 
counter-current  condenser,  illustrating 
the  principles  of  a  barometric  jet  con- 
denser.    The  cooling  water  enters  the 
upper  part  of  the  condensing  chamber 
A  through  pipe  N  and  falls  in  cascades, 

as  shown  in  the  figure,  to  tail  pipe  B,  from  which  it  flows  by  gravity  to 
the  hot  well.  The  exhaust  steam  enters  chamber  A  through  pipe  D, 

*  The  author  has  been  informed  that  the  word  "  Barometric  "  in  connection  with 
jet  condensers  is  the  registered  trade  mark  of  the  Alberger  Condenser  Company. 


FIG.  208. 


Weiss  Counter-Current 
Condenser. 


412 


STEAM  POWER  PLANT  ENGINEERING 


and,  coming  in  contact  with  the  cold-water  spray,  is  condensed.  The 
air  is  exhausted  from  the  top  of  the  condenser  by  a  dry  vacuum  pump 
through  pipe  F.  In  flowing  to  the  pump  the  air  passes  upwards 
through  the  water  spray  and  its  temperature  is  lowered  to  that  of  the 
injection  water,  thereby  reducing  the  volume  to  be  exhausted.  Any 
moisture  passing  over  with  the  air  is  separated  at  G  before  reaching 
the  air  pump,  and  flows  out  through  the  small  barometric  tube  H. 
The  cooling  water  is  forced  to  the  condenser  chamber  through  pipe  N 
by  any  positive  displacement  pump,  the  actual  head  pumped  against 


Fio.  209.     Section  Through  Condensing  Chamber,  Alberger  Barometric  Condenser. 

being  the  difference  between  the  total  height  and  that  of  a  column  of 
water  corresponding  to  the  degree  of  vacuum  in  the  condenser.  The 
main  barometric  tube  or  tail  pipe  B  through  which  the  water  is  dis- 
charged is  34  feet  or  more  in  length  and  is  provided  with  a  foot  valve  C. 
The  counter-current  principle  permits  a  much  higher  temperature  of 
hot  well  for  the  same  degree  of  vacuum  than  does  the  parallel  current, 
a  hot-well  temperature  of  120  degrees  and  a  vacuum  of  27  inches  being 
readily  maintained.  A  small  pipe  K  connecting  the  main  condenser 


CONDENSERS  413 

with  the  small  barometric  tube  H  insures  at  all  times  a  sufficient 
quantity  of  water  in  the  small  auxiliary  hot  well  to  seal  the  tube. 
The  water  from  this  auxiliary  hot  well  flows  over  a  weir,  as  indicated, 
into  a  counter-weighted  bucket  M,  the  latter  having  a  hole  in  the  bot- 
tom which  allows  the  normal  flow  to  escape.  But  in  case  a  sudden 
heavy  overload  is  thrown  on  the  engines,  and  the  adjustment  is  for  a 
light  load,  the  temperature  of  the  discharge  will  reach  the  boiling  point 
and  an  abnormal  quantity  of  water  will  flow  down  the  small  barometric 
tube.  This  will  cause  the  water  to  flow  into  the  bucket  much  faster 
than  the  opening  in  the  bottom  can  dispose  of  it ;  as  a  result  the  bucket 
will  increase  in  weight  and  will  open  up  a  free-air  valve  L  which 
reduces  the  vacuum  two  or  three  inches  and  raises  the  boiling  point 
without  "  dropping  "  the  vacuum  entirely.  E  is  the  atmospheric  relief 
valve. 

Fig.  209  shows  a  section  through  the  condensing  chamber  of  an 
Alberger  barometric  condenser.  In  principles  of  operation  the  con- 
denser is  similar  to  the  Weiss,  but  differs  considerably  in  details. 
Exhaust  steam  enters  at  A  and  divides  into  two  streams,  one  flowing 
directly  to  the  inner  chamber  Z),  the  other  through  the  annular  space  E. 
Cooling  water  enters  through  B  and  is  broken  up  into  a  fine  spray 
by  the  serrated  cone  F,  which  is  hung  upon  a  long  spring,  thus  auto- 
matically adjusting  itself  to  the  quantity  of  water  entering  the  con- 
denser. After  condensing  the  exhaust  steam  in  the  inner  cylinder  the 
partly  heated  spray  of  cooling  water  in  falling  is  brought  in  contact 
with  the  exhaust  steam  which  enters  through  the  annular  space. 
This  process  permits  of  a  high  hot-well  temperature  without  affecting 
the  degree  of  vacuum.  The  air  which  is  not  entrained  by  the  cooling 
water  and  carried  down  the  tail  pipe  collects  under  the  spray  cone  F 
and  ascends  through  the  tubular  support  of  the  cone  into  the  air 
cooler.  This  air  cooler  is  simply  a  small  chamber  in  which  the  non- 
condensable  gases  are  cooled  by  a  small  portion  of  the  circulating 
water  before  they  are  withdrawn  by  the  air  pump.  The  circulating 
water  used  for  the  purpose  is  forced  into  the  cooling  chamber 
through  pipe  K  and  falls  through  serrated  openings  in  the  bottom 
to  the  condenser  proper.  The  air  enters  the  chamber  through 
these  same  openings,  and  is  withdrawn  by  the  air  pump.  Surround- 
ing the  cooler  is  a  separating  space  of  large  capacity  to  allow  the 
subsidence  of  any  entrained  moisture  before  the  air  reaches  the  vacuum 
pump. 

Fig.  236  shows  a  typical  installation  of  an  Alberger  condenser  in 
connection  with  a  cooling  tower,  and  Fig.  226  that  of  a  Weiss  condenser 
in  the  Northwestern  Elevated  R.  R.  Power  Station,  Chicago. 


414 


STEAM  POWER  PLANT  ENGINEERING 


Fig.   210  shows   a  section  through   the   condensing  chamber  of 
Worthington  barometric  condenser.     The  drawing  is  self-explicit. 


HAND    WHEEL 


ENGINE 


TO  TAIL,  PIPE 

Fro.  210.     Section  Through  Condensing  Chamber,  Worthington  Barometric  Condenser. 

Fig.  211  shows  a  section  through  a  Tomlinson  barometric  condenser. 
The  air  pump  instead  of  discharging  into  the  atmosphere  is  made  to 
deliver  into  the  tail  pipe  where  the  vacuum  is  still  sufficient  to  support 
the  column  of  water  below  the  point  of  delivery.  The  effect  produced 
is  that  of  a  two-stage  air  pump,  the  tail  pipe  becoming  the  second 
stage.  Suitable  by-pass  valves  enable  the  air  pump  to  be  discharged 
into  the  atmosphere  or  to  be  cut  out  entirely.  (Power,  February, 
1907,  p.  94.) 

Fig.  21  la  shows  the  application  of  a  centrifugal  pump  to  the  tail 
pipe  of  a  barometric  condenser.  This  permits  of  a  very  short  tail  pipe, 
as  the  pump  takes  the  place  of  the  barometric  column. 

;  Counter-Current  Condensers  :  Am.  Elecn.,  Feb.,  1905,  p.  81;  Power,  March,  1905, 
p.  182,  Jan.,  1906,  p.  44;  Engr.  U.S.,  Jan.,  1906,  p.  58;  Hausbrand,  Evaporating 
and  Condensing  Apparatus,  Chapter  XX;  Bulletin  No.  6,  Heisler  Mfg.  Co.,  St. 
Marys,  Ohio.  The  Barometric  Condenser:  Power,  Jan.,  1907,  p.  1. 


CONDENSERS 


415 


INJECTION 


FIG.  211.     Tomlinson  Type  B  Barometric  Condenser. 


416 


STEAM  POWER  PLANT  ENGINEERING 


DISCHARGC 


As  previously  outlined,  surface  condensers  may  be  divided  into  three 
general  classes,  (a)  water  cooled,  (6)  air  cooled,  and  (c)  evaporative. 

219.  Water-Cooled  Surface 
Condensers.  —  Water  -  cooled 
surface  condensers  are  by  far 
the  most  extensive  in  use  and 
only  occasionally  are  the  con- 
ditions such  as  to  warrant  the 
installation  of  the  other  class. 
They  are  ordinarily  classified 
as  (1)  single-flow,  (2)  double- 
flow,  and  (3)  multi-flow. 

Fig.  212  shows  a  sectional 
FlGV21-ilaii-  Ceiltrifligal  pumP  Applied  to  the     elevation  through  a  Baragwa- 

Tail  Pipe  of  a  Barometric  Condenser. 

nath  vertical  condenser,  illus- 
trating the  single- 
flow  type.  It  con- 
sists essentially  of 
a  cast  -  iron  shell 
provided  with  two 
heads,  into  which  a 
number  of  one-inch  EXHAUST 
brass  tubes  are  ex- 
panded. Exhaust 
steam  fills  the  shell 
and  flows  around 
and  between  the 
tubes,  while  the 
cooling  water  is 
caused  to  circulate 
through  the  tubes 
by  means  of  a 
circulating  pump. 
The  steam  is  con- 
densed by  contact 
with  the  tubes  and 
drops  to  the  bottom 
tube  sheet,  from 
which  it  is  exhausted 
by  the  air  pump. 
The  circulating 


STEAM 


WATER 
INLET 


TO    AIR   PUMP 


FIG.  212.     Baragwanath  Surface  Condenser. 


water  flows  through  the  tubes  in  one  direction  only,  hence  the  name 
"  single  flow."     To  allow  for  the  unequal  expansion  of  shell  and  tubes 


CONDENSERS 


417 


418 


STEAM  POWER  PLANT  ENGINEERING 


the  two  halves  of  the  shell  are  provided  with  slightly  thinner  plates 
flanged  outward,  the  flanges  being  bolted  together  with  a  spacing  ring 
between  them.  This  joint  gives  to  the  shell,  in  the  direction  of  its 
length,  a  certain  amount  of  elasticity  which  is  sufficient  to  allow  for  the 
greatest  possible  elongation  of  the  tubes  without  straining  the  tube  ends 
and  causing  leakage. 

Fig.  213  shows  a  section  through*  a  Wheeler  admiralty  surface  con- 
denser mounted  on  a  combined  air  and  circulating  pump,  illustrating 
the  typical  "  double-flow  "  surface  condenser.  The  condenser  proper 
consists  of  a  ribbed  cast-iron  chamber  of  rectangular  section  fitted 
with  a  number  of  small  seamless  drawn  brass  tubes  through  which 
the  cooling  water  is  forced  by  suitable  means.  The  exhaust  steam 
enters  at  the  top  and  is  prevented  from  impinging  directly  against  the 
tubes  by  baffle  plates,  which  serve  also  to  distribute  the  steam  more 
evenly  over  the  cooling  surface.  The  steam  in  passing  between  the 
tubes  is  condensed,  and  falls  to  the  bottom  of  the  chamber,  from  which 
it  is  removed,  together  with  the  entrained  air,  by  a  vacuum  pump. 
The  water  chamber  between  the  tube  sheet  and  the  head  is  divided  into 
two  compartments,  as  shown  in  the  illustration,  the  partition  being  so 
arranged  that  the  water  flows  first  through  the  lower  set  of  tubes  and 
then  through  the  upper  set  in  the  opposite  direction.  Thus  the  tem- 
perature of  the  cooling  water  increases  as  it  rises,  and  reaches  a  maximum 
where  the  exhaust  steam  enters;  Condensation  begins  as  soon  as  the 
vapor  enters  the  condenser,  and  the  surfaces  of  the  tubes  are  at  once 
covered  with  a  thin  film  of  water  flowing  downwards  from  tube  to 
tube. 


FIG.  214.     Surface  Condenser,  C.  H.  Wheeler  &  Co. 

Fig.  214  gives  the  details  of  a  C.  H.  Wheeler  &  Co.'s  high- 
vacuum  surface  condenser.  The  condensing  chamber  is  of  the  series- 
parallel  type  in  which  the  water  enters  the  top  group  of  tubes,  then 
passes  to  the  middle  section  and  finally  through  the  bottom  section. 


CONDENSERS 


419 


Connecting  chambers  are  provided  at  the  ends  of  the  shell  as  illustrated. 
This  construction  of  water  chamber  keeps  the  condenser  completely 
filled  with  cooling  water  at  all  times.  The  inlet  is  at  the  bottom  but  the 
water  is  carried  up  through  the  annular  chamber  to  the  top  of  the  tubes. 


FIG.  215.     Weighton  Multi-Flow  Surface  Condenser. 

Fig.  215  shows  a  section  through  a  multi-flow  surface  condenser 
designed  by  Professor  R.  L.  Weighton.  The  condenser  has  three 
compartments  separated  by  two  diaphragms  inclined  to  the  hori- 
zontal. Each  compartment  is  fitted  with  a  number  of  brass  tubes 
three-fourths  inch  in  diameter  by  four  feet  in  length,  spaced  one 
and  one-eighth  inches  between  centers.  The  cooling  water  circulates 
through  the  tubes  five  times,  giving  an  effective  length  of  20  feet.  The 
notable  features  of  the  condenser  are  abolishment  of  steam  space,  and 
compartment  drainage  of  condensed  steam.  Mere  passages  of  such  shape 


420  STEAM  POWER  PLANT  ENGINEERING 

and  section  as  will  insure  distribution  of  the  steam  over  the  entire  surface 
are  used  instead  of  the  large  steam  space  usually  associated  with  surface 
condensers.  Each  compartment  is  separately  drained  to  the  air  pump, 
so  that  the  surfaces  in  the  lower  compartments  are  unimpeded  in  their 
condensing  action  by  the  condensed  steam  from  the  upper  compartments 
flowing  over  them.  Referring  to  Fig.  215:  Exhaust  steam  enters  the 
condenser  at  A  and  flows  toward  the  hot  well  H.  The  greater  part 
of  the  steam  is  condensed  in  the  first  section  of  the  condenser,  and  the 
condensation  is  drained  directly  to  the  hot  well.  The  balance  of  the 
condensation  takes  place  in  the  remaining  sections,  the  condensed  steam 
being  withdrawn  from  each  section.  The  wet-air  pump  withdraws  the 
condensed  steam  and  non-condensable  gases  through  opening  P.  Cool- 
ing water  enters  at  /  and  leaves  at  0.  An  exhaustive  series  of  tests  on 
a  condenser  of  this  type  credit  it  with  a  much  higher  efficiency  than 
the  ordinary  single  or  double-flow  apparatus.  (See  "  The  Efficiency  of 
Surface  Condensers,"  Proc.  Institute  of  Naval  Architects,  March,  1906; 
also  Engineer,  London,  April  27,  1906.) 

220.  Cooling  Water,  Surface  Condensers.  —  The  amount  of  cooling 
water  required  per  pound  of  steam  in  a  surface  condenser  is  dependent 
upon  the  vacuum,  the  temperature  of  the  condensed  steam,  and  the 
range  in  temperature  of  the  cooling  water;  it  may  be  closely  approxi- 
mated from  the  formula 

w-  x~TTj+T32> 

•*•  2~   20 

where 

A.  =  total  heat  of  the  exhaust  steam  above  32  degrees  F. 
T!=  temperature  of  the  condensed  steam. 
TQ=  temperature  of  the  injection  water. 
T2=  temperature  of  the  discharge  water. 

W=  pounds  of  injection  water  necessary  to  condense  one  pound 
of  steam. 

Example  :  Required  the  quantity  of  cooling  water  necessary  to  con- 
dense one  pound  of  steam  under  the  following  conditions:  Initial  tem- 
perature of  the  cooling  water  60  degrees  F.;  final  temperature  100 
degrees  F.;  vacuum  26  inches,  referred  to  30-inch  barometer.  Here 
X  =  1120  B.T.U.,  TQ=  60,  T2  =  100. 


That  is,  the  ratio  of  cooling  water  to  condensed  steam  is  26.0  to  1. 
In  turbine  practice  where  vacua  as  high  as  one-half  pound  absolute  are 


CONDENSERS 


421 


obtained,  the  ratio  of  cooling  water  to  condensed  steam  is  nearly  twice 
this  quantity.  For  example,  if  a  vacuum  of  28.92  inches  is  desired 
with  the  barometer  at  29.92  and  the  range  of  the  circulating  water  tem- 
perature is  70  to  50  degrees  and  the  temperature  of  the  hot  well  80 
degrees,  the  ratio  will  be 

1106-80  +  32 


W  *= 


70-50 


=  52.9. 


In  determining  the  amount  of  cooling  water  it  is  well  to  bear  in 
mind  that  in  the  ordinary  condenser  of  the  single  or  double-flow  type 


140 
130 
120 
110 
100 
90 
80 
70 
CO 
50 
40 

^^. 

A   New  Type  with  Cores  and  Spray  and 
Dry  Air  Pump 
B   New  Type  without  Cores  and  Spray 
Ordinary  Pump 

. 

"'~—  •—  ». 
-^ 

^^ 
-^ 

\ 

>> 

^s 

s^ 

C    Old  Type,  Ordinary  Pump 

- 

^ 

*^ 

X 

x 

^A 

*^ 

\. 

^ 

8s 

\ 

x 

I" 

^^ 

^ 

\ 

\ 

0. 

^ 

x 

\ 

\ 

i 

X 

x 

\ 

\ 

\<i 

Relation  between  Hot-Well  Temperature 
and  Vacuum  in  Surface  Condensers 

•  ^ 

s 

\ 

\ 

\ 

\ 

Vacuum  Referred  to  30  Inch  Barometer 
FIG.  216. 

the  temperature  of  the  condensed  steam  will  be  from  10  to  20  degrees 
lower  than  that  corresponding  to  the  degree  of  vacuum  in  the  con- 
denser, and  that  the  temperature  of  the  condensing  water  at  the  dis- 
charge point  will  be  from  5  to  10  degrees  lower  than  the  temperature 
due  to  the  vacuum. 

With  well-designed  condensers  of  the  multi-flow  type  the  temper- 
ature of  the  hot  well  may  be  from  3  to  5  degrees  higher  than  the  tem- 
perature due  to  the  vacuum,  and  the  temperature  of  the  condensing 
water  at  the  discharge  point  may  be  equal  to  or  slightly  higher  than 
that  due  to  the  vacuum.  (Proc.  Inst.  of  Naval  Arch.,  March,  1906.) 
(See  Fig.  216.) 

221.  Extent  of  Water-Cooling  Surface.  —  Theoretically  the  opera- 
tion of  a  surface  condenser  is  divided  into  two  periods,  (1)  the  period 


422  STEAM  POWER  PLANT  ENGINEERING 

of  condensation  during  which  the  heat  of  vaporization  at  the  observed 
pressure  is  removed  and  (2)  the  period  of  cooling  during  which  the 
temperature  of  the  condensed  steam  is  reduced.  In  order  to  determine 
accurately  the  extent  of  cooling  surface  it  would  be  necessary  to  cal- 
culate the  heat  transmission  for  each  of  the  two  periods.  In  practice, 
however,  it  is  assumed  that  condensation  and  cooling  take  place  simul- 
taneously, and  that  the  mean  temperature  difference  is  a  direct  function 
of  the  temperature  corresponding  to  the  exhaust  steam  in  the  con- 
denser and  that  of  the  condensed  steam  and  cooling  water.  The  error 
in  these  assumptions  has  only  a  slight  influence  on  the  estimation  of  the 
cooling  surface  and  is  entirely  lost  sight  of  in  the  liberal  factor  allowed 
in  practice. 

Let  S  =  cooling  surface  in  square  feet. 

A.  =  total  heat  above  32  degrees  of  the  exhaust^iteam  at  con- 

denser pressure. 

TQ=  initial  temperature  of  the  circulating  water. 
T2=  final  temperature  of  the  circulating  water. 
TI  =  final  temperature  of  the  condensed  steam. 
Ti  =  temperature  of  the  exhaust  steam  at  condenser  pressure. 
U  =  coefficient  of  heat  transmission,  B.T.U.  per  hour,  per  degree 

difference  in  temperature,  per  square  foot  of  cooling  sur- 

face. 

d  =  mean  difference  in  temperature  between  Ts  and  T2,  and  ro. 
W  =  weight  of  condensed  steam  per  hour. 

d=     T?~  Tn       (see  equation  (118),  Chap.  XII); 


and  since  the  heat  absorbed  by  the  cooling  water  is  equal  to  the  heat 
given  up  by  the  steam, 

SUd  =  W{\-(Tl-32}}.  (91) 


(92) 


Whitham  ("Steam-Engine  Design,"  p.  283)  uses  the  arithmetic  mean 

rp       i      rp 

d  =  Ts  --  i-5  —  2  instead  of  the  mean  as  determined  from  (118). 

Equation  118  is  based  on  the  assumption  that  the  fluid  on  each  side 
of  the  tube  is  homogeneous,  which  is  far  from  being  true  in  the  case  of 
the  air-steam  mixture  in  a  condenser,  and  for  this  reason  many  designers 
prefer  to  use  the  simpler  arithmetic  formula. 

The  coefficient  of  heat  transfer,  [7,  as  used  in  above  equations  refers 


CONDENSERS 


423 


tc  the  mean  or  average  values  for  the  entire  surface  since  the  actual  heat 
transmission  varies  widely  for  different  parts  of  the  condenser;  thus 
the  actual  value  of  U  varies  from  over  1000  in  the  first  few  rows  of  the 
tubes  (where  the  steam  comes  directly  into  contact  with  the  cooling 
surface)  to  less  than  50  in  the  bottom  row  (where  the  tubes  are  practi- 


FIG.  216a.     Application  of  Weighton  Dry-tube  Surface  Condenser  to 
Vertical  Marine  Engine. 

cally  submerged  in  water  of  condensation)  and  to  3  or  less  for  the  tubes 
surrounded  only  by  air. 

Prof.  Josse  of  the  Royal  Technical  School,  Charlottenburg,  after  an 
exhaustive  investigation  of  the  subject  found  that  the  actual  value  of 
U  varied  with 

(1)  The  material,  thickness,  shape  and  cleanliness  of  the  tubes. 

(2)  The  velocity  of  the  water  through  the  tubes. 


424 


STEAM  POWER   PLANT  ENGINEERING 


(3)  The  velocity  of  the  steam  against  the  tubes. 

(4)  The  percentage  of  air  in  the  steam  surrounding  the  tubes. 

(5)  The  extent  of  submersion  of  the  steam  side  of  the  tubes. 

Some  of  the  results  of  his  investigations  are  shown  in  Figs.  216,  b, 

c,  and  d.     See  also  Power  and 
Engr.,  Feb.  2,  1909. 

The  effect  of  thickness,  ma- 
terial, etc.,  of  condenser  tubes  is 
so  small  in  the  ultimate  result 
and  the  choice  and  arrangement 
are  so  largely  determined  by 
practical  consideration  that  they 
may  be  neglected. 

The  value  of  U  increases  ap- 
proximately as  the  square  root 
of  the  velocity  of  the  water  flow- 
ing within  the  tube,  so  that  in- 
crease in  water  velocity  effects  a 
substantial  increase  in  the  heat 
transmission  ;  but  the  resistance 
encountered  by  the  circulating 
water  increases  as  the  square  of 
the  velocity,  and  the  power  con- 
sumed in  pumping  the  water  in- 
creases as  the  third  power  of  the 
velocity,  so  that  a  point  is  soon 
reached  where  the  gain  on  the 
one  hand  may  be  offset  by  the 
loss  on  the  other. 

A  study  of  a  number  of  instal- 
lations gave 
Old-style  surface  condenser, 

V  =  30  to  240  ft.  per  min.,  average  90. 
Modern  dry  tube  surface  condenser, 

V  =  120  to  360  ft.  per  min.,  average  240. 

From  the  curves  in  Figs.  126b  and  126c  it  will  be  seen  that  air  is  an 
excellent  heat-insulating  material;  hence,  the  greater  the  amount  of 
air  entrained  with  the  steam  the  lower  will  be  the  coefficient  of  heat 
transmission.  The  necessity  of  removing  the  air  as  fast  as  it  accumu- 
lates is  at  once  apparent. 

In  the  older  types  of  surface  condensers  the  water  of  condensation 
from  the  upper  rows  of  tubes  is  permitted  to  fall  on  the  rows  immedi- 


1234567 

Rate  of  Flow  of  Cooling  Water-  Feet  per  Second 

FIG.  216b. 


CONDENSERS 


425 


14 


Heat  Transference  for  Air 


10        20        30        40        50        60 
Mean  Rate  of  Flow  of  Air  —Feet  per  Second 

FlG.  21 6c. 


70 


ately  below,  the  water  increasing  in  volume  as  it  passes  the  successive 

banks  of  tubes  until  it  completely  envelops  them.     The  coefficient  U 

varies  from  1000  or  more  in  the 

upper  row  to  less  than  50  in  the 

lower,  giving  a  mean  value  of 

approximately  250  to  350  for  the 

entire  surface.      In  estimating 

the  extent  of  cooling  surface  for 

a    condenser   of   this    type    an 

average  figure  for   plain  brass 

tubes  with  water  velocities  of  50 

to  100  feet  per  minute  is  U  =  250. 

For  a  velocity  of  100  to  240  feet 

per  minute  U  may  be  taken  50per 

cent  greater  than  these  figures. 

When  the  tubes  are  clean  a  much 

higher  value  may  be  taken,  but 

a  liberal  factor  is  usually  allowed 

for  possible  variation  in  the  con- 
dition of  operation. 

In  the  modern  dry-tube  surface  condenser,  designed  along  the  lines 

of  the  one  described  in  paragraph  221,  in  which  the  water  of  condensa- 
tion is  withdrawn  as  rapidly  as 
it  is  formed,  mean  values  of 
U  =  800  to  900  are  not  unusual. 
In  estimating  the  extent  of 
cooling  surface  for  condensers 
of  this  type  an  average  value 
of  U  is  600  with  water  veloci- 
ties of  4  to  5  feet  per  second. 

Example:  Standard  Type  of 
Surface  Condenser :  —  Required 
the  number  of  square  feet  of 
cooling  surface  perl. H. P.  neces- 
sary to  condense  the  steam  from 
an  engine  operating  under  the 
following  conditions:  Engine 

uses   20   pounds   of  steam  per  I.H.P.-hour,   vacuum   26  inches 

barometer  at  30;  temperature  of  cooling  water  at  60  degrees. 
Here  ^  =  1115  and  T8  =  126  (from  steam  tables), 
T0  =  60, 
Tl  =  T9-1Q=  116. 


Heat  Transference  for  Air 


25          20          15         10  5 

Air  Pressure*- Inches  of  Mercury 

FIG.  216d. 


with 


426  STEAM   POWER  PLANT  ENGINEERING 

In  this  type  of  condenser  average  practice  gives  a  temperature 
difference  of  approximately  10  degrees  between  the  temperature  of  the 
hot  well  and  that  corresponding  to  the  degree  of  vacuum. 

T2  =  Ts  -  15  =  101. 

Any  value  may  be  fixed  upon  for  T2  greater  than  T0  and  less  than  Ts. 
The  nearer  T2  is  to  T0  the  greater  must  be  the  quantity  of  circulating 
water  per  unit  of  time  for  a  given  rate  of  condensation.  On  the  other 
hand,  the  nearer  T2  is  to  Ts  the  less  is  the  mean  temperature  difference 
d  and  hence  the  greater  must  be  the  cooling  surface  for  a  given  coeffi- 
cient of  heat  transmission.  When  water  is  cheap  and  the  head  pumped 
against  is  small  T.,  should  be  given  a  lower  value  than  when  water  is 
costly  and  the  discharge  head  is  large.  Average  engine  practice,  with 
conditions  as  stated,  gives  T2  a  value  of  approximately  15  degrees  less 
than  that  corresponding  to  the  degree  of  vacuum. 
The  logarithmic  mean  is,  equation  (118), 

101  -  60 

*-,       1-          =  42A 
lofe 


126  -  101 
The  arithmetic  mean  gives 

d  =  126  _  eo^ioi  =  45.5.  | 

Substitute  the  value  of  d  in  equation  (92)  and  assume  U    =  250,  the 
figure  commonly  used  for  this  type  of  condenser. 
_      20(1115-  116  +  32)  _ 

250X42.4" 
or  say  two  square  feet  per  I.H.P.  of  engine. 

Surface  condensers  of  this  type  are  ordinarily  rated  on  a  basis  of  two 
square  feet  per  I.H.P. 

Example:  Dry-tube  Multi-flow  Surface  Condenser:  —  Required  the 
number  of  square  feet  of  cooling  surface  per  kilowatt  necessary  to 
condense  the  steam  from  a  steam  turbine  operating  under  the  following 
conditions:  Turbine  uses  15  pounds  of  steam  per  kw.-hour;  vacuum 
28.5  inches,  referred  to  30-inch  barometer;  temperature  of  cooling 
water  70  degrees. 

Here  ^  =  .9  X  1100  =  990. 

The  total  heat  of  dry  steam  corresponding  to  an  absolute  pressure 
of  1.5  inches  is  1100,  but  in  the  case  of  high  vacuum  turbine  practice 
the  steam  entering  the  condenser  is  far  from  being  dry,  the  quality 
varying  from  0.80  to  0.95,  depending  upon  the  quality  of  the  steam  at 
admission.  An  average  figure  is  0.9. 

Ts  =  92,         T0  =  70,         Tl  =  Ts  -  4  =  88. 


CONDENSERS 


427 


In   this   type   of   condenser   the   hot -well   temperature   varies  from 
T1  =  Ts  to  7\  =  Ts  -  8. 

T2  =  Ts  -  5  =  87. 

In  this  type  T2  varies  from  T2  =  Ts  to  T2  =  Ts  -  10. 

87-70 


d  = 


log 


90-70 

92  -  87 


=  11.5. 


Arithmetic  mean  gives  d  =  92  -     — = —  =  13-5. 

Substitute  the  value  of  d  in  equation  (92)  and  assume  U  =  600,  the 
figure  commonly  used  for  this  type  of  condenser. 
15  (990  -  88  +  32) 


600  X  11.5 


=  2.02, 


or  say  2  square  feet  per  kw.  of  generator.  There  is  no  standard 
rating  of  surface  condenser  for  steam  turbine  work  because  of  the  wide 
variation  in  operating  conditions.  A  study  of  a  number  of  modern 
installations  gives 

1  to  2.5  sq.  ft.  per  kw.  for  large  turbo-generators  using 
dry  tube  surface  condensers. 

2.5  to  4  sq.  ft.  per  kw.  for  small  turbo-generators  using 

standard  surface  condensers. 

Professor  Weighton  found  from  his  experiments  that  a  surface  con- 
denser constructed  on  the  lines  of  the  one  described  in  paragraph  221  in 
conjunction  with  dry-air  pumps,  was  capable  of  condensing  20  pounds 
of  steam  per  square  foot  of  surface  per  hour  and  maintained  a  vacuum 
of  28^  inches  (referred  to  a  30-inch  barometer),  and  this  with  a  cooling- 
water  consumption  of  24  pounds  per  pound  of  condensed  steam;  with 
an  inlet  temperature  of  50  degrees  F.  a  condensation  of  35  pounds  of 
steam  per  hour  per  square  foot  of  cooling  surface  was  effected  at  a 
ratio  of  28  pounds  of  cooling  water  per  pound  of  steam,  the  vacuum 
remaining  28J  inches.  See  Fig.  216.  (Engineering  Record,  May  19,  1906, 
p.  615.) 

EXAMPLES    OF    MODERN    CONDENSER    PROPORTIONS. 


Name  of  Station. 

Size  of  Turbo- 
Generators. 

Sq.  Ft.  of 
Condenser 
Surface. 

Sq.  Ft.  of 
Surface 
per  Kw. 

Commonwealth  Edison  Co.: 
Northwest  Station  

20,000 

32,000 

1   60 

Quarry  Street 

14  000 

25  000 

1   79 

Fisk  Street 

12  000 

25  000 

2  08 

*59th  St.,  Interborough,  N.  Y  

15,000 

25,000 

1.67 

Metropolitan  St.  Ry.,  Kansas  City  

10,000 

22,000 

2.20 

*  Combined  Engine  and  Low-pressure  Turbine. 

SURFACE  CONDENSER  AIR  PUMPS.  —  See  paragraphs  284-291. 


428 


STEAM   POWER  PLANT  ENGINEERING 


TABLE   52. 

SQUARE    FEET   OP  COOLING    SURFACE   NECESSARY   TO   CONDENSE   AND   COOL 
ONE    POUND    OF    STEAM    PER   MINUTE. 

(Barometer  29.92.) 


Temp, 
of  In- 
jection 
Water. 

Vacuum  24". 
Temperature  of  Steam  141°. 

Temp, 
of  In- 
jection 
Water. 

Vacuum  25". 
Temperature  of  Steam  134°. 

Temperature  of  Hot  Well. 

Temperature  of  Hot  Well. 

110 

115 

120 

125 

130 

105 

110 

115 

120 

125 

40 
50 
60 
70 
80 
90 

3.1 
3.3 
3.7 
3.9 
4.1 
4.4 

3.3 
3.5 
3.8 
4.1 
4.4 
4.8 

3.5 
3.7 
4.0 
4.3 

4.7 
5.1 

3.7 
4.0 
4.2 
4.6 
5.0 
5.5 

4.0 
4.3 
4.6 
5.0 
5.5 
6.0 

40 
50 
60 
70 
80 
90 

3.3 
3.5 
3.8 
4.1 
4.5 
5.1 

3.5 
3.7 
4.0 
4.4 

4.8 
5.4 

3.7 
4.0 
4.3 

4.7 
5.2 

5.8 

4.0 
4.3 
d.« 

5.1 
5.6 
6.2 

4.4 
4.7 
5.1 
5.5 
6.1 
6.9 

Temp, 
of  In- 
jection 
Water. 

Vacuum  26". 
Temperature  of  Steam  125°. 

Temp, 
of  In- 
jection 
Water. 

Vacuum  27". 
Temperature  of  Steam  114°. 

Temperature  of  Hot  Well. 

Temperature  of  Hot  Well. 

100 

105 

110 

115 

90 

95 

100 

105 

40 
50 
60 
70 
80 

3.6 
3.8 
4.2 
4.6 
5.1 

3.9 
4.2 
4.6 
5.1 
5.6 

4.2 
4.6 
5.0 
5.4 
6.1 

4.6 
5.0 
5.4 
6.0 
6.7 



40 
50 
60 

70 

80 

4.1 
4.4 
4.8 
5.4 

4.4 
4.7 
5.2 
5.8 
6.7 

4.7 
5.1 
5.G 
6.3 
7.2 

5.1 
5.6 
6.2 
7.0 
8.0 



Temp, 
of  In- 
jection 
Water. 

Vacuum  28". 
Temperature  of  Steam  100°. 

Temp, 
of  In- 
ection 
Water. 

Vacuum  29". 
Temperature  of  Steam  77°. 

Temperature  of  Hot  Well. 

Temperature  of  Hot  Well. 

75 

80 

85 

90 

60 

65 

70 

40 
50 
60 
70 

4.6 
5.0 
5.5 

4.9 
5.4 
6.1 

5.3 

5.8 
6.6 

7.7 

5.8 
6.4 
7.3 
8.5 



35 
40 
45 
50 

6.3 
6.8 
7.4 

6.9 
7.4 
8.0 
8.9 

7.7 
8.3 
9.0 
9.9 



Final  temperature  of  injection  water  assumed  to  be  10  degrees  lower  than  that  of  the  hot-well. 

222.  Dry-Air  Surface  Condensers  (Forced  Circulation).  —  Where 
water  is  very  scarce  and  the  feed  supply  is  reclaimed  by  condensing  the 
exhaust  steam,  water-cooled  condensers  may  be  prohibitive  in  cost  of 
operation,  even  when  combined  with  cooling  tower  or  other  water-cool- 


CONDENSERS 


429 


ing  device,  since  the  latter  involves  a  loss  of  water  approximately 
equivalent  to  the  amount  of  steam  condensed,  due  to  evaporation. 

Under  these  conditions  air  cooling  has  been  successfully  adopted. 
In  the  city  of  Kalgoorlie,  West  Australia,  an  electric  station  of  2000- 
horse-power  capacity  is  equipped  with  air-cooled  surface  condensers. 
The  condensers  have  been  in  use  five  years  (1906),  and  have  given 
excellent  service  with  very  little  expense  and  maintenance.  The  con- 
denser consists  of  a  large  number  of  narrow  chambers  constructed  of 
thin  corrugated  sheet-steel  plates  spaced  \  inch  between  centers.  Each 
chamber  has  1345  square  inches  of  cooling  surface.  Fifty-one  of 
these  chambers  are  grouped  into  a  compartment  and  15  compartments 
constitute  a  section.  Each  section  is  equipped  with  three  motor-driven 
fans  7  feet  in  diameter  and  running  normally  at  320  r.p.m.  In  all 
there  are  six  sections,  giving  a  total  cooling  surface  of  45,000  square 
feet.  The  steam  consumption  of  the  main  engines  is  16  to  16.5  pounds 
per  I.H.P.  hour  at  rated  load.  At  full  load  the  fans  require  130  kilo- 
watts, or  approximately  10  per  cent  of  the  station  output.  The  average 
vacuum  obtained  is  about  18  inches  throughout  the  year  and  ranges 
from  0  inches  on  very  hot  days  to  22  inches  in  cooler  weather.  The 
following  figures,  based  on  actual  observation,  show  the  effect  of  tem- 
perature of  the  external  air  on  the  vacuum  when  condensing  32,000 
pounds  of  steam  per  hour  (the  rated  capacity  of  the  condenser) . 


Temperature  Ex- 
ternal Air, 
Degrees  F. 

Vacuum,  Inches 
(referred  to  30-Inch 
Barometer)  . 

Temperature  Ex- 
ternal Air, 
Degrees  F. 

Vacuum,  Inches 
(referred  to  30-Inch 
Barometer). 

42.8 
50 
60.8 
68 

78.8 

22 
21.2 
20 
18.4 
16 

96.8 
100.4 
107.6 
113 

9.6 
7.6 
3.6 

0 

Air-Cooled  Surface  Condensers  :  Engineering  News,  Oct.,  1902,  p.  271 ;  ibid.,  Vol. 
49,  p.  203. 

223.  Quantity  of  Air  for  Cooling  (Dry-Air  Condenser).  —  The  volume 
of  air,  under  atmospheric  conditions,  necessary  to  condense  steam  to 
any  given  temperature  may  be  determined  as  follows: 

Let  X  =  total  heat,  above  32  degrees  F.  of  the  steam  at  condenser 

pressure. 

Ts  =  temperature  of  the  vapor  in  the  condenser. 
Tl  =  temperature  of  the  condensed  steam. 
t  =  temperature  of  the  air  entering  condenser. 
^  =  temperature  of  the  air  leaving  condenser. 
V  =  volume  of  air  in  cubic  feet  necessary  to  condense  and  cool 
one  pound  of  steam. 


480  STEAM  POWER  PLANT  ENGINEERING 

B  =  specific  weight  of  air  under  atmospheric  conditions. 
C  =  specific  heat  of  air  under  atmospheric  conditions. 
d  =  mean  temperature  difference  between  the  air  and  steam. 
S  =  cooling  surface  in  square  feet. 

U  =  coefficient  of  heat  transmission,  B.T.U.  per  square  foot  per 
degree  difference  in  temperature  per  hour. 

Since  the  heat  absorbed  by  the  air  must  be  equal  to  the  heat  given 
up  by  the  steam,  neglecting  radiation  we  have 

VBC^-t)  =  A-7\  +  32,  (93) 

from  which 

X-ri+32 
BC  (t.-t) 

For  practical  purposes  C  may  be  taken  as  the  specific  heat  of  dry  air, 
the  error  due  to  this  assumption  being  negligible  even  if  the  air  is 
saturated  with  moisture. 

Example  :  How  many  cubic  feet  of  air  are  necessary  to  condense  and 
cool  one  pound  of  steam  under  the  following  conditions  :  Vacuum  20 
inches;  temperature  of  entering  air,  leaving  air,  and  condensed  steam, 
60,  110,  and  140  degrees  F.  respectively? 

Here  X  =  1131  (from  steam  tables). 

^  =  110,  T1  =  140,  t  =  60,  C  =  0.2377,  B  =  0.075. 

Substituting  these  values  in  equation  (94), 
1131  —  140  4-  32 


0.075X0.2377(110-60)  °f  M 

condense  one  pound  of  steam  under  the  given  conditions. 

The  proper  area  of  cooling  surface  depends  upon  the  value  of  the 
coefficient  of  heat  transmission,  which  varies  with  the  velocity  and 
humidity  of  the  air  and  character  of  the  cooling  surface.  Accurate 
data  are  not  available  on  this  point. 

A  few  experiments  made  at  the  Armour  Institute  of  Technology 
gave  values  of  U  =  10  to  25  B.T.U.  per  hour,  per  square  foot,  per 
degree  difference  in  temperature  for  air  velocities  of  500  to  4000  feet 
per  minute  for  corrugated  steel  sheeting  J  inch  thick.  Hence,  sub- 
stituting in  equations  (94)  and  (92)  we  get,  for  the  above  example. 
S  =  1.5  square  feet  of  cooling  surface  per  pound  of  steam  condensed 
per  hour  for  air  velocity  of  4000  feet  per  minute,  and  S  =  3.7  square 
feet  for  a  velocity  of  500  feet  per  minute. 

224:.  Saturated-Air  Surface  Condensers  (Natural  Draft).  —  Fig.  217 
shows  vertical  and  horizontal  sections  of  a  Fennel  saturated-air  surface 


CONDENSERS 


431 


condenser.  The  apparatus  consists  of  an  upright  cylindrical  shell 
containing  a  number  of  vertical  4-inch  steel  tubes  through  which  air  is 
drawn  by  natural  draft.  A  centrifugal  pump  circulates  about  one- 
half  gallon  of  water  per  horse  power  per  minute  from  a  cistern  below  the 
condenser.  The  water  flowing  over  the  upper  tube  sheet  and  then 
descending  the  tubes  by  gravity  forms  a  film  over  their  entire  interior 
surface. 


Horizontal  Section.  Section  on  AB. 

FIG.  217.    Fennel  Saturated-Air  Surface  Condenser. 


The  condensing  action  is  as  follows:  The  current  of  exhaust  steam 
entering  the  side  of  the  shell  at  A  is  caused  by  suitable  baffle  plates  to 
circulate  among  the  tubes,  and  in  condensing  gives  up  its  latent  heat 
to  the  water  film,  which  wholly  or  partially  evaporates,  saturating  the 
ascending  current  of  air  at  its  own  temperature.  The  upward  current 
of  hot  vapor-laden  air  carries  off  the  heat  into  the  atmosphere.  The 
cooling  water  which  is  not  evaporated  and  lost  to  the  atmosphere  falls 
into  the  cistern  below  to  be  again  taken  up  by  the  circulating  pump, 
the  water  level  in  the  cistern  being  kept  constant  by  a  float  governing 
a  valve  on  the  supply  pipe.  The  non-condensable  gases  collect  at  C, 
where  they  are  removed  by  the  dry-air  pump,  while  the  condensed  steam 
is  drawn  off  from  the  bottom  tube  sheet  by  the  vacuum  pump  and 
discharged  into  the  hot  well.  An  excellent  feature  of  this  device  is 
that  the  film  of  water  on  the  cooling  surface  is  secured  without  inter- 
ference with  the  ascending  air  currents  and  also  without  the  use  of 
sprays  through  small  orifices  likely  to  become  clogged  with  rust  or 
sediment.  Where  the  recovery  of  the  condensed  steam  is  essential  and 
a  high  vacuum  of  secondary  importance,  condensers  of  this  type  have 
proved  to  be  good  investments  on  account  of  the  low  first  cost. 


432 


STEAM  POWER  PLANT  ENGINEERING 


Table  53  gives  the  results  of  a  test  of  a  condenser  of  this  type,  taking 
steam  from  a  30  x  58  x  48  engine  running  at  45  r.p.m.  (Power,  December, 
1903,  p.  672;  West.  Elect,  May  19,  1900,  p.  323.) 

TABLE  53. 
TEST    OF    FENNEL    SATURATED-AIR    SURFACE    CONDENSER. 

Duration  of  trial 9  hours 

Average  steam  pressure  at  engine  by  gauge 139.8  pounds 

Average  vacuum,  mercury  column  17.5  inches 

Average  temperature  in  condenser 123.7  degrees  F. 

Average  temperature  of  circulating  water 116.4        do 

Average  temperature  of  city  water 52  do 

Average  temperature  of  outside  air 62  do 

Average  temperature  of  saturated  air 106  do 

Average  draft  in  stack  of  condenser 1.1  inches 

Average  humidity  of  outside  air 67  per  cent 

Average  amount  of  steam  condensed  per  hour 7950  pounds 

Average  amount  of  circulating  water  used  per  hour 114,660  pounds 

Average  amount  of  city  water  used  per  hour 3462  pounds 

Pounds  of  city  water  per  pound  of  steam 2.3 

Pounds  of  circulating  water  per  pound  of  steam 14.4 

Average  horse  power  of  engine 569.7 

Steam,  pounds  per  I.H.P.  per  hour 13.95 

Horse  power  required  to  run  air  pumps 10 . 5 

Horse  power  required  to  run  circulating  pumps 3.0 

Condensing  surface,  square  feet 3900 

Pounds  of  steam  condensed  per  square  foot  surface  per  hour 2038 

Barometer 28.58  inches 

Vapor  tension  corresponding  to  123.7  degrees 3.82  inches 

Per  cent  of  main  engine  steam  used  by  auxiliaries 2.38 

Fig.  218  illustrates  the  Fennel 
"  flask "  type  of  atmospheric 
condenser.  The  exhaust  steam 
enters  below  and  follows  the  zig- 
zag course  bounded  by  the  inter- 
nal stay  channels,  condensing  as 
it  goes  and  driving  before  it  the 
non-condensable  gases  to  the  out- 
let at  the  top.  The  condensed 
steam  gravitates  to  the  bottom 
and  thence  to  the  hot  well.  The 
top  of  the  flask  is  trough  shaped 
and  causes  the  cooling  water  to 
flow  down  the  sides  of  the  flask 
in  a  thin  stream.  The  portion 

of  the  cooling  water  not  evaporated  collects  at  the  bottom  of  the  flask 

and  flows  to  the  cooling-water  reservoir. 


FIG.  218.   Fennel  Flask  Type  of  Saturated- 
Air  Surface  Condenser. 


CONDENSERS  433 

225.  Evaporative  Surface  Condensers.  —  An  evaporative  surface  con- 
denser consists  of  a  number  of  copper,  brass,  wrought-  or  cast-iron 
tubes  arranged  horizontally  or  vertically  and  connected  to  manifolds 
or  chambers  at  each  end.  The  exhaust  steam  passes  through  the 
tubes  and  a  thin  film  of  water  is  allowed  to  flow  over  the  external 
surfaces.  The  cooling  effect  is  brought  about  by  the  evaporation  of 
part  of  the  circulating  water,  and  the  general  principle  of  operation 
is  the  same  as  that  of  the  saturated-air  condenser  described  above. 
Evaporation  is  sometimes  hastened  by  constructing  a  flue  over  the 
tubes,  thereby  creating  a  natural  draft,  or  by  means  of  fans.  With 
horizontal  cast-iron  tubes  and  natural  draft,  vacua  from  23  to  27  inches 
are  readily  maintained  with  a  cooling  surface  of  approximately  eight- 
tenths  square  foot  per  pound  of  steam  condensed  per  hour.  With 
vertical  brass  tubes  and  fan  draft  8  pounds  of  steam  per  hour  per 
square  foot  of  cooling  surface  is  not  an  unusual  figure.  The  amount 
of  cooling  water  evaporated  per  pound  of  steam  varies  from  eight- 
tenths  to  one  pound,  depending  upon  the  draft.  The  power  necessary  to 
operate  the  pumps  and  fans  varies  from  1  to  4  per  cent  of  the  total 
output  of  the  plant.  For  an  interesting  discussion  of  evaporative 
condensers  the  reader  is  referred  to  the  admirable  article  by  Oldham 
in  the  Proceedings  of  the  Institute  of  Mechanical  Engineers,  1899,  and 
reproduced  as  a  serial  in  Engineering  (London),  April  28  to  June  30, 
1899.  The  following  test  of  a  vertical  cast  iron  tube  evaporative 
surface  condenser  (Table  54)  will  give  some  idea  of  the  performance 
of  this  type  of  condenser.  This  condenser  consisted  of  two  rows  of 
4-inch  vertical  cast-iron  pipes  connected  at  the  top  by  U  bends  and  at 
the  bottom  by  cast-iron  manifolds.  A  perforated  iron  trough  dis- 
tributes the  water  over  the  center  of  the  bend  and  causes  it  to  flow  in 
a  thin  stream  over  the  surface  of  the  tubes.  A  wet-air  pump  is  used 
for  withdrawing  the  condensed  steam  and  air.  No  fan  is  used  for 
hastening  evaporation.* 

Evaporative  Condensers:  Engr.,Lond.,May  5,  1899,  pp.  432,  442,  447;  Engineering, 
May  19,  1899,  p.  661,  June  2,  1899,  p.  721,  June  30,  1899,  p.  861;  Trans.  A.S.M.E. 
14-696;  Power,  Sept.,  1904,  p.  542;  Prac.  Engr.  U.S.,  June,  1910,  p.  346. 

226.  Location  and  Arrangement  of  Condensers.  —  In  the  modern 
power  house  one  sees  two  general  arrangements  of  condensers  and 
auxiliaries : 

1.  The  independent  or  subdivided  system,  in  which  each  engine  or 
turbine  is  provided  with  its  own  condenser,  air  and  circulating  pumps. 

2.  The  central  system,  in  which  the  condensers  and  auxiliaries  are 
grouped  together.     Ordinarily  one  condenser  suffices  for  all  engines. 

*  See  end  of  paragraph  236  for  evaporated  surface  condenser  calculations. 


434 


STEAM  POWER  PLANT  ENGINEERING 


TABLE   54. 

TEST  OF  A  CAST-IRON,  VERTICAL-TUBE,  EVAPORATIVE   SURFACE    CONDENSER, 

NATURAL  DRAFT. 


Date  

Sept   12 

Sept    13 

Weather 

Wet 

Fine 

Barometer 

29  8 

29  5 

Temperature  of  air 

1 

60 

Cooling  surface  external 

272 

272 

Duration  of  trial,  minutes 

99 

115 

Weight  of  steam  condensed,  pounds  . 

800 

800 

Boiler  pressure  

60 

60 

Weight  of  water  in  circulation  

1830 

1830 

Weight  of  fresh  water  added  

600 

640 

Vacuum  in  condenser  

23  36 

24  1 

Initial  temperature  of  circulating  water  
Final  temperature  of  circulating  water 

117.5 
128  4 

113.9 
125 

Temperature  of  "  make  up  "  water 

58 

58 

Temperature  of  water  in  hot  well 

136  5 

131  8 

Weight  of  steam  condensed  per  hour,  pounds  .  .  . 
Weight  of  water  circulated  per  hour,  pounds  
Weight  of  "  make-up  "  water  added  per  hour.  .  . 
Weight  of  steam  condensed  per  square  foot  of 
cooling  surface  per  hour 

485 
6786 
364 

1  8 

427 
? 
334 

1  54 

Weight  of  "  make-up  "  water  per  pound  of  steam 
condensed  pounds              

0  75 

0  80 

226a.  The  Independent  System. — The  condenser  is  usually  placed 
close  to  and  below  the  engine  so  that  all  condensation  may  gravitate 
into  it.  Figs.  219  and  221  show  an  application  of  this  system  with 
jet  condensers.  Here  each  condenser  receives  its  supply  of  cooling 
water  from  a  main  injection  pipe  and  discharges  into  a  main  overflow 
pipe.  The  exhaust  pipe  leading  to  the  condenser  is  by-passed  through 
a  suitable  atmospheric  relief  valve  to  a  main  free  exhaust  header  so 
that  the  engine  may  operate  non-condensing  in  case  the  vacuum  breaks 
or  the  condenser  is  cut  out.  The  chief  feature  of  this  arrangement  is 
its  flexibility,  as  each  unit  is  complete  in  itself  and  independent  of 
the  others.  By  far  the  greater  number  of  central  stations  are  equipped 
with  independent  condensers. 

Occasionally  a  jet  condenser  is  located  on  the  same  level  with  the 
engine  or  even  above  it,  Fig.  222,  but  such  a  location  should  be  avoided 
if  possible,  as  it  usually  necessitates  a  larger  number  of  bends  and 
joints  in  the  exhaust  pipes  than  the  basement  arrangement,  and 
increases  the  possibility  of  air  leakage.  If  the  exhaust  pipe  does  not 
drain  directly  into  the  condenser,  the  lowest  point  in  the  piping  should 
always  be  provided  with  a  drip  which  should  be  opened  when  the 
engine  is  shut  down,  as  condensation  and  leakage  are  apt  to  fill  the 
pipe  with  water  if  the  engine  stands  for  any  length  of  time.  The  end 


CONDENSERS 


435 


FIG.  219.    Jet  Condenser  located  below  Engine-Room  Floor. 


FIG.  220.    Surface  Condenser  located  below  Engine-Room  Floor. 


436 


FIG.  221.     Surface  Condenser,  Installed  in  the  Suction  Line  of  a  Pumping  Engine. 


FIG.  222.    Jet  Condenser  located  above  Engine-Room  Floor. 


FIG   223.     Typical  Arrangement.  Westinghouse-Leblanc 
Condenser  and  Curtis  Turbine 


CONDENSERS 


487 


d 


438 


STEAM   POWER  PLANT  ENGINEERING 


of  the  drip  should  be  connected  so  that  water  cannot  be  drawn  back 
through  the  drip  pipe  and  into  the  engine  cylinder.  The  length  of 
exhaust  pipe  and  particularly  the  number  of  bends  between  engine  and 
condenser  should  be  kept  as  small  as  possible,  otherwise  the  engine 
may  not  derive  the  full  benefit  of  the  vacuum  in  the  condenser.  A 
case  is  recorded  where  the  exhaust  piping  and  appurtenances  in  con- 
nection with  a  5000-horse-power  engine  caused  a  drop  of  several  inches 
in  vacuum  between  condenser  and  exhaust  opening  of  the  low-pressure 
cylinder.  (National  Engineer,  December,  1906,  p.  10.)  The  wet-air 
pump  must  always  be  located  below  the  condenser  chamber  so  that  the 
condensation  may  gravitate  to  it. 


6   CX.HAUST.                              .^J                  ,y.                        \          \  d 

V      FUTUA£  3O~£XHAUS7^"^k U-1 V-1 ^  •" 

d T~i  II 


--fi— --== ==?}- 


/M%K/ 
flj 


FIG.  225.    Plan  of  Piping  for  Engine  and  Condenser,  Des  Moines  City  Ry.  Co. 


Fig.  252  shows  a  surface  condenser  installation  in  connection  with  a 
vacuum  or  primary  heater. 

Fig.  236  shows  an  application  of  a  barometric  condenser  to  a 
vertical  engine  installation. 

Fig.  220  shows  the  arrangement  of  a  surface  condenser  with  com- 
bined air  and  circulating  pump  in  connection  with  a  horizontal  cross 


CONDENSERS  439 

compound  engine.  The  condenser  and  appurtenances  are  placed  below 
the  engine,  thereby  permitting  the  condenser  to  be  closely  connected 
to  the  engine. 

Fig.  221  shows  the  arrangement  of  a  surface  condenser  in  connec- 
tion with  a  pumping  engine.  The  condenser  is  placed  in  series  with 
the  pump  suction. 

227.  Central  Systems.  —  In  the  central  condensing  systems  the  con- 
denser is  located  at  any  convenient  point  and  the  exhaust  from  all  the 
engines  piped  to  it.  Any  arrangement  of  condenser  and  auxiliary 
machinery  may  be  adopted  which  will  favor  the  lowest  cost  of  installa- 
tion and  expense  of  operation.  Except  where  continuity  of  opera- 
tion is  absolutely  essential,  only  one  circulating  pump  and  one  air  pump 
are  installed.  This  reduces  the  number  of  auxiliary  pumps  and  appliances 
to  a  minimum,  with  a  consequent  decrease  in  first  cost  and  maintenance. 
With  properly  designed  exhaust  piping  the  condenser  may  be  located 
at  a  considerable  distance  from  the  engine  without  undue  loss  of  vacuum. 
At  the  Cambria  Steel  Works,  Johnstown,  Pa.,  the  maximum  drop 
between  condenser  and  engine  is  only  three-quarters  of  an  inch  and  the 
distance  between  them  is  about  1000  feet. 

Central  condensers  have  found  great  favor  in  power  plants  in  which 
the  individual  units  are  subjected  to  extreme  variations  in  load,  as  in 
rolling  mills.  At  the  works  of  the  Illinois  Steel  Company,  South  Chicago, 
111.,  one  condenser  takes  care  of  the  steam  from  15,000  horse  power  of 
engines  in  the  rail  mill,  and  another  condenses  the  steam  from  the 
15,000  horse  power  of  engines  in  the  Bessemer  steel  mill.  A  notable 
installation  of  this  system  in  connection  with  street-railway  work  is 
in  the  power  house  of  the  Northwestern  Elevated  Company,  Chicago, 
where  a  single  condenser  takes  care  of  the  exhaust  steam  of  five  engines, 
11,000  horse  power  in  all.  Fig.  226  shows  the  general  arrangement  of 
this  installation. 

For  a  comparison  of  the  advantages  and  disadvantages  of  the  inde- 
pendent and  central  systems  see  Engineering  Magazine,  October,  1900, 
p.  56,  Engineering,  London,  June  23,  1899,  p.  615,  and  Engineering,  July 
17,  1903. 

Centralization  of  Steam-Condensing  Plant:  Eng.  Mag.,  Oct.,  1900,  p.  56;  Iron  Age, 
Jan.  7,  1904;  Revue  Technique,  Feb.  25,  1903. 

Five  Thousand  H.P.  Surface-Condensing  Plant :  Engr.,  Lond.,  May  23,  1903. 

Aurora  &  Elgin  R.R.  Condenser  Plant :  Engr.  Rec.,  Vol.  47,  p.  153. 

Condensing  Apparatus  of  Manhattan  Elevated  Power  Plant:  Power,  Aug.,  1903, 
p.  411. 

Interborough  R.R.  Condenser  Plant :   St.  Ry.  Jour.,  Oct.  8,  1904. 

New  York  Rapid  Transit  Condenser  Plant :   Power,  June,  1903,  p.  283. 

Worthington  Surface  Condensers  for  Metropolitan  Power  Station :  Power,  June, 
1901,  p.  15. 


440 


STEAM  POWER  PLANT  ENGINEERING 


CONDENSERS 


441 


228.  High-vacuum  Systems.  —  The  average  reciprocating  engine 
gives  its  best  commercial  economy  at  a  vacuum  of  approximately  26 
inches  (referred  to  a  30"  barometer),  and  the  ordinary  standard  jet  or 
surface  condenser  has  been  designed  to  meet  this  requirement.  At 
the  time  of  the  introduction  of  the  steam  turbine  it  was  discovered  that 
a  very  high  vacuum  would  improve  turbine  economies  to  an  extent 
hitherto  impossible  when  applied  to  reciprocating  engines.  This  con- 
dition naturally  created  an  era  of  development  among  the  condenser 


FIG.  226a.     Condenser  Installation,  Quincy  Point  Power  Plant  of  the 
Old  Colony  Street  Railway  Company. 


designers.  It  became  evident  at  once  that  the  old  types  that  were 
capable  of  creating  a  26"  or  27"  vacuum  would  require  considerable 
modification  to  maintain  a  vacuum  of  28"  or  29".  The  principal 
improvement  adopted  by  practically  all  manufacturers  has  been  to 
apply  a  separate  dry  vacuum  pump  for  the  removal  of  air  and  non- 
condensable  vapors. 

Surface  Condensers.  —  Fig.  227  shows  the  arrangement  advocated 
by  the  H.  R.  Worthington  Company.  The  equipment  comprises  a 
surface  condenser,  a  steam-driven  centrifugal  pump  for  circulating 
the  cooling  water,  a  steam-driven  rotative  dry-air  pump,  and  a  motor- 


442 


STEAM   POWER  PLANT  ENGINEERING 


CONDENSERS 


443 


driven  centrifugal  hot-well  pump.  The  surface  condenser  is  piped 
direct  to  the  turbine  exhaust,  only  a  corrugated  copper  expansion  joint 
and  a  tee  intervening.  A  tubular  water  vapor  cooler,  which  is  in  reality 
a  small  surface  condenser,  is  inserted  in  the  circulating  water  line 
between  the  pump  suction  and  condenser,  and  serves  to  arrest  all  the 
condensable  vapor  and  thus  reduces  the  volume  to  be  handled  by  the 
air  pump.  All  condensation,  including  that  from  the  air  cooler,  collects 
in  the  hot  well,  from  which  it  is  pumped  by  a  motor-driven  circulating 
pump  direct  to  heater  or  boiler.  Cooling  water  is  handled  by  a  cen- 
trifugal pump  having  both  suction  and  delivery  pipes  water-sealed,  so 
that  the  work  done  by  the  pump  is  virtually  that  of  overcoming  the 


YIG.  228.     High- Vacuum  System,  C.  H.  Wheeler  Co. 


fluid  friction  in  the  condenser  and  piping.  All  valves  and  stuffing 
boxes  are  water-sealed  to  prevent  any  possible  leakage  of  air,  and  the 
condenser  pump  cylinder  is  especially  designed  to  avoid  vapor  binding. 
This  makes  it  possible  to  maintain  a  vacuum  of  one-half  pound  absolute 
with  cooling  water  at  60  degrees  F.  In  the  high-vacuum  condenser 
installation  of  the  Chicago  Edison  Company  the  hot-well  pump,  dry-air 
pump,  and  the  circulating  pump  are  direct  connected  to  a  single- 
cylinder  Corliss  engine. 

Fig.  228  shows  the  general  arrangement  of  the  C.  H.  Wheeler  Com- 
pany's high-vacuum  condensing  outfit.  The  condensing  chamber  is 
shown  in  section  in  Fig.  214  and  is  described  in  paragraph  219.  The 
wet-air  pump  is  illustrated  in  Fig.  287  and  is  described  in  paragraph  288. 
No  dry-air  pump  is  needed,  and  the  makers  guarantee  a  vacuum  within 


444 


STEAM   POWER  PLANT  ENGINEERING 


two  inches  of  absolute  under  full-load  conditions  of  operating  steam 
turbines. 

Fig.  229  shows  a  section  through  a  Parsons  "  vacuum  augmenter  " 
for  increasing  the  vacuum  in  a  surface  condenser.  A  pipe  is  led  from 
the  bottom  of  the  main  condenser  to  an  auxiliary  or  augmenter  having 
about  one-twentieth  of  the  cooling  surface  of  the  main  condenser.  At 
the  point  indicated  a  small  steam  jet  is  provided  which  acts  as  an  ejector 
and  draws  out  the  air  and  vapor  from  the  condenser  and  delivers  it  to 
the  air  pump.  The  water  seal  prevents  the  air  and  vapor  from  returning 
to  the  condenser.  With  this  arrangement,  according  to  tests  conducted 
by  Mr.  Parsons,  if  there  is  a  vacuum  of  27^  or  28  inches  in  the  condenser, 
there  may  be  only  26  at  the  air  pump,  which,  therefore,  may  be  of  small 


FIG.  229.    Parsons  Vacuum  Augmenter. 


size,  the  jet  compressing  the  air  and  vapor  from  the  condenser  to  about 
one-half  of  its  original  volume.  The  steam  jet  uses  about  one  and  one- 
half  per  cent  of  the  steam  used  by  the  turbine  at  full  load. 

Jet  Condensers.  —  Fig.  229a  gives  the  general  details  of  a  Westing- 
house- Leblanc  multi-jet  condenser  which,  under  commercial  conditions, 
has  realized  vacua  within  99  per  cent  of  the  ideal.  The  most  striking 
feature  of  this  system  lies  in  its  compactness  and  simplicity,  a  1500- 
kw  equipment  being  less  than  9  feet  in  height.  Referring  to  Fig. 
229a,  exhaust  steam  enters  the  condenser  chamber  at  the  upper  left- 
hand  opening  and  meets  the  cooling  water  as  it  is  forced  through  spray 
nozzle  C.  The  condensed  steam  and  injection  water  fall  to  the  bottom 
of  the  condenser  and  are  removed  by  centrifugal  pump  M.  The  non- 
condensable  vapors  are  withdrawn  by  valveless  rotary  air  pumps  P, 
through  suction  opening  0.  Referring  to  section  N-N  through  the 


CONDENSERS 


445 


air  pump  it  will  be  seen  that  this  pump  consists  primarily  of  a  reverse 
Pelton  turbine  wheel  in  conjunction  with  an  ejector.  Sealing  water  is 
introduced  through  the  branch  indicated  by  dotted  outline,  into  the 
central  chamber  G,  from  which  it  passes  through  port  H.  It  is  then 
caught  up  by  the  blades  P  of  the  Pelton  wheel,  which  is  rotated  at  a 
suitable  speed,  and  ejected  into  the  discharge  cone  in  the  form  of  thin 


SECTION  N.-N. 
THROUGH  AIR  PUMP. 


SECTION    M.-M. 
THROUGH    WATER  PUMP. 


FIG.  229a.     Westinghouse-Leblanc  Multi-jet  High-vacuum  Condenser  System. 


sheets  having  a  high  velocity.  These  sheets  of  water  meet  the  sides 
of  the  discharge  cone  and  thus  form  a  series  of  water  pistons,  each  of 
which  entraps  a  small  pocket  of  air  and  forces  it  out  against  the  atmos- 
pheric pressure.  In  passing  through  the  air  pump  the  sealing  water 
receives  practically  no  increase  in  temperature,  hence  the  same  water 
may  be  used  over  and  over  again.  The  air  pump  rotor  and  main  pump 
runner  are  enclosed  in  a  common  casing  mounted  on  the  same  shaft. 


446 


STEAM  POWER  PLANT   ENGINEERING 


This  arrangement  makes  the  plant  very  compact  and  requires  the  use 
of  only  one  motor  to  drive  both  pumps.  There  is  a  clear  passage  throurh 
the  condenser  and  pump,  so  that  should  the  pump  stop  for  any  reason 


Exhaust  from  Turbin« 


FIG.  229b.     Tomlinsoii  Type  C  High-vacuum  Jet  Condenser. 

air  rushes  into  the  condenser  through  the  air  pump  and  immediately 
breaks  the  vacuum.  In  starting  up  the  condenser,  steam  is  turned  into 
auxiliary  nozzle  L,  section  N-N,  for  a  few  moments,  thus  creating  suf- 
ficient vacuum  to  start  the  regular  flow  of  water  through  the  air  pump. 


FIG.  229c.  Section  through  Condensing  Chamber  of  Kbrting  Multi-jet  Condenser. 
Chamber  Capable  of  Maintaining  a  Vacuum  of  95  per  Cent  of  the  Ideal  without 
the  Use  of  Air  Pumps. 

The  pumps  require  from  H  to  3  per  cent  of  the  power  generated  by 
the  main  engines.  Fig.  223  shows  an  application  of  a  Westinghouse- 
Leblanc  condenser  to  a  Curtis  turbine. 


CONDENSERS  447 

229.  Power  Consumption  of  Condenser  Auxiliaries.  —  In  estimating 
the  cost  of  producing  vacua  with  the  different  types  of  auxiliaries, 
steam  driven,  electrically  driven,  or  belted,  the  power  consumption  is 
most  conveniently  expressed  in  terms  of  the  equivalent  heat  consumption 
of  the  auxiliary  in  question  and  not  the  indicated  or  developed  power. 
For  example,  suppose  a  power  plant  has  a  number  of  1200-1.  H.  P. 
engines  direct  connected  to  800-kilowatt  generators  and  that  the  engines 
use  20  pounds  of  steam  per  I.H.P.  hour  at  rated  load;  furthermore 
suppose  the  engine  driving  the  air  pump  (jet  condenser)  to  indicate 
24  horse  power.  Now,  it  is  manifestly  incorrect  to  say  that  the  power 

consumption  of  the  air  pump  is  equivalent  to  -  =  2  per  cent  of  the 


main  engine  power  unless  the  engine  driving  the  air  pump  uses  20 
pounds  of  steam  per  I.H.P.  As  a  matter  of  fact  the  small  engine  proba- 
bly uses  30  to  40  pounds  or  more  of  steam  per  I.H.P.  hour,  and  the  true 
power  consumption  is 

24  X  30 


1200  X  20 


3  per  cent,  or  more. 


If  the  exhaust  steam  is  piped  to  the  condenser,  then  all  of  this  3  per 
cent  or  more  should  be  charged  against  the  condenser;  if  the  steam  is 
piped  to  a  heater,  then  only  the  difference  between  the  heat  enter- 
ing the  small  engine  and  that  given  up  to  the  feed  water  should  be 
charged  against  it.  For  example,  suppose  the  engine  in  the  preceding 
examples  uses  30  pounds  of  steam  per  I.H.P.  hour  when  running 
condensing  and  40  pounds  when  operating  non-condensing.  Let 
the  initial  steam  pressure  be  150  pounds  and  feed-water  temperature 
120  degrees  F.  when  the  air  pump  is  running  condensing.  If  the  boiler 
feed  is  not  taken  from  the  hot  well,  the  heat  in  the  exhaust  steam  is 
lost  so  far  as  the  economy  of  the  plant  is  concerned,  and  the  heat  con- 
sumption per  I.H.P.  hour  is  30(1193.6  -  (120  -  32)}=  33,168  B.T.U. 
This  represents  the  cost,  in  heat  units,  of  producing  the  vacuum,  and 
is  equivalent  to  3  per  cent  of  the  main  engine  output. 

If  the  air  pump  runs  non-condensing  and  the  exhaust  steam  is  piped 
to  the  heater,  each  pound  of  exhaust  steam  gives  up  approximately 
950  B.T.U.  per  hour  to  the  feed  water  and  the  temperature  of  the 
latter  is  raised  from  120  to  180  degrees  F.  The  heat  entering  the  air 
pump  is  40(1193.6 -(120 -32)}  =44,224  B.T.U.  per  I.H.P.  hour. 
But  40  X  950  =  38,000  B.T.U.  are  returned  to  the  feed  water.  Hence 
44,424  —  38,000  =  6224  is  the  net  heat  consumption  of  the  air  pumps 
per  I.H.P.  hour.  This  corresponds  to  approximately  0.55  per  cent  of 
the  main  engine  output. 


448 


STEAM  POWER  PLANT  ENGINEERING 


In  the  preceding  example  suppose  the  air  pump  to  be  motor  driven 
and  that  it  requires  20  electrical  horse  power  per  hour.     This  will  be 

20 

the  equivalent  of   — =  26.2  I.H.P.  of  the  main  engine  on  the 

0.85  X  0.90 

assumption  that  the  efficiency  of  the  small  motor  is  85  per  cent  and  that  of 
the  engine  and  generator  combined  90  per  cent.  The  power  required  by 
the  air  pump  will  be  26.2  -4-  1200  =  2.2  per  cent  of  the  total  output. 


UL 


RELATION  OP  POWER  CONSUMPTION 
OF  AUXILIARIES  TO  STATION 

OUTPUT 
Citizens  Light,  Heat  and  Power  Co. 

'Johnsto-wn,  P.O. 

Estimated 

_  Auxiliary  Input  #  of  Turbine  Output  I.H.P. 
"  "  «  Generator     "      E.H.P. 


700 


FIG.  230. 


In  practice  the  auxiliaries  use  the  equivalent  of  from  1  to  15  per  cent 
of  the  main  engine  or  turbine  steam,  depending  upon  the  size  of  the  plant, 
character  and  number  of  auxiliaries,  and  the  conditions  of  operation. 


POWER  CONSUMPTION  OF  AUXILIARIES 
2000  K.  W.  Curtis  Turbine 
La.  Purchase  Exposition 


1500  Load  JK^W.         2000 

Fro.  231. 


B5UU 


Table  55  gives  the  power  consumption  of  the  condenser  auxiliaries 
in  a  number  of  installations.  Fig.  230  shows  the  relation  between  the 
power  consumption  of  the  auxiliaries  and  the  total  output  of  the 


CONDENSERS 


449 


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450  STEAM  POWER  PLANT  ENGINEERING 

station  at  different  loads  for  a  Parsons  steam  turbine  installation,  and 
Fig.  231  shows  a  similar  relation  for  a  2000-kilowatt  Curtis  turbine. 
(J.  R.  Bibbins,  Power,  January,  1905.) 

Steam  vs.  Electric  Auxiliaries  :  Engr.  U.S.,  1902,  p.  113;  Power,  Feb.,  1905,  p.  90, 
Sept.,  1906,  p.  502;  St.  Ry.  Review,  March,  1898,  p.  184,  July,  1899,  p.  458. 
Centralized  Control  of  Auxiliaries  :   Engr.  U.S.,  Nov.  1,  1906,  p.  782. 

230.  Cost  of  Condensers.  —  The  following  figures  give  an  idea  of 
the  relative  costs  of  the  different  types  of  condensers  and  auxiliaries 
for  a  1000-I.H.P.  plant  using  20  pounds  of  steam  per  I.H.P.  hour  at 
rated  load,  or  a  total  of  20,000  pounds  per  hour.  Vacuum  to  be 
maintained,  26  inches,  unless  otherwise  stated;  temperature  of  cooling 
water,  70  degrees  F.;  hot-well  temperature,  105  to  120  degrees  F.;  dis- 
tance between  engine  exhaust  opening  and  mean  level  of  intake  well, 
10  feet. 

Siphon  Condensers. 

1  16"  siphon  condenser  with  6"  centrifugal  pump  driven  by  6"  by  6* 

vertical  engine $800 

Jet  Condensers. 

1    14"  by  22*  by  24"  jet  condenser  with  single  horizontal  direct-acting 

pump „ 1335 

1    16"  by  24"  by  18"  jet    condenser   with    single   vertical   direct-acting 

pump 1620 

1    14"  by  24"  by  18"  jet  condenser  with  single  vertical  fly  wheel  vacuum 

pump 1770 

1  12"  by  17"  by  22"  by  25"  jet  condenser,  single  horizontal  direct- 
acting  compound  pump . .  „ 2200 

Barometric  Condensers. 

1  barometric  condenser;  10"  by  16"  by  12"  horizontal  single-cylinder 
rotative  dry-air  pump;  8"  horizontal  volute  centrifugal  pump  direct 
connected  to  23-horse-power  high-speed  engine 2500 

1  barometric  condenser;  16"  by  16"  dry-air  pump  direct  connected  to 
9"  by  16"  steam  engine;  positive  rotary  pump,  for  circulating 

cooling  water,  belted  to  above  engine 4300 

Surface  Condensers. 

1   surface  condenser,    1025  square  feet  cooling    surface,  mounted   over 

1\"  by  14"  by  14"  by  12"  combined  air  and  circulating  pump 2100 

1  surface  condenser,  1025  square  feet  cooling  surface,  with  7^"  by  12" 
by  12"  horizontal  air  pump,  direct  acting,  and  6"  centrifugal  pump 
driven  by  5"  by  5"  engine 2300 

1  surface  condenser,  1025  square  feet  cooling  surface;  5"  by  12"  by 
10"  Edwards  single-cylinder  air  pump  and  6"  centrifugal  pump 
driven  by  a  5"  by  5"  engine ;  maximum  28",  referred  to  30"  barometer  . .  2850 

1  surface  condenser,  1025  square  feet  cooling  surface;  6"  by  8"  rotative 
dry-air  pump ;  6"  by  6"  Edwards  wet-air  pump  and  6"  centrifugal 
pump  driven  by  5"  by  5"  engine;  maximum  vacuum  29",  referred  to 
30"  barometer  (temp,  cooling  water  50  deg.  F.) 3500 


CONDENSERS 


451 


In  general  the  cost  of  complete  condensing  equipments  installed  and 
ready  for  operation  will  approximate  as  follows: 

Cost  per  Kilowatt  of  Main 
Generating  Unit. 

Siphon  condensers  without  air  pump $2.00  to  $  3.00 

Jet  condensers 3.00  to      4.50 

Barometric  condensers  with  dry-air  pump 4.00  to      6.00 

Surface  condensers  for  26-inch  vacuum  . . . . 3.50  to      5.00 

High- vacuum  surface  condensers 3.50  to    10.00 

The  curve  in  Fig.  232  shows  the  relative  costs  of  complete  surface 
condensing  plants  for  steam  turbines  to  maintain  the  vacua  indicated. 

It  will  be  noted  how  much  more 
expensive  a  high-vacuum  plant  is 
than  one  designed  for  moderate 
vacua.  Thus  a  27- inch  plant 
costs  25  per  cent  more  than  a 
26 -inch  plant,  and  a  28.5 -inch 
plant  costs  twice  as  much.  (J.  R. 
Bibbins,  Power,  January,  1905.) 

26  n  I  \Jf\  I  I  I  I  I  I  I  I  I  I  f T"is>,  T ?8.  e,tci  I  I        The  real  cost  of  a^  condensing 

plant,  however,  is  not  limited  to 
the  cost  of  condensing  auxiliaries 
and  piping,  but  should  include 
all  other  costs  necessitated  by 
the  use  of  the  condensing  plant,  including  cost  of  extra  building  space, 
foundations  and  the  like,  and  the  attending  fixed  charges. 

231.  Most  Economical  Vacuum.*  —  The  load  factor,  or  the  ratio 
of  the  actual  yearly  load  to  the  rated  yearly  capacity,  has  a  marked 
influence  on  the  degree  of  vacuum  best  suited  for  a  given  installation, 
since  the  fixed  charges  go  on  whether  the  plant  is  running  or  not,  while 
the  gain  due  to  the  higher  vacuum  is  realized  only  when  the  engines 
are  operating.  The  higher  the  load  factor  the  greater  is  the  amount  of 
power  produced,,  the  longer  does  the  apparatus  operate  at  best  effi- 
ciency, the  lower  the  ratio  of  fixed  charges  to  total  operating  expenses, 
and  consequently  the  lower  the  cost  of  power  per  unit. 

The  load  factor  for  electric-lighting  stations  is  invariably  low  and 
seldom  exceeds  25  per  cent,  with  an  average  not  far  from  18  per  cent. 
In  street-railway  work  it  is  higher  and  averages  about  30  per  cent.  In 
manufacturing  plants  the  load  factor  varies  considerably,  but  as  a  rule 
is  somewhat  higher  than  in  either  of  the  above  cases.  Tables  56  and 
57  (Power,  December,  1906,  p.  769)  show  the  most  economical  vacua 
for  different  load  factors  for  plants  of  1000  kilowatts  capacity  with 
*  See  also,  Elec'n,  Lond.,  Jan.  14,  1910. 


1    1 

\ 

1 

1 

RELATIVE  COST  OP  HIGH  \A 
Condens'ng  Apparatus 

LCUTJM 

g 

3.  — 

> 

**** 

^^ 

— 

:=» 

^  — 

gg 

ifi 

°f" 

^ 

^ 

^ 

S 

S 

Cunil 
£urfa 
Dry  i 
pircu 
,Hotl 
Pipin 

rising 
3e  Condeasere 
iir  Pumps, 
latiug  Pumps, 
Yell  Pumps, 
g,  Valves  etc. 

/ 

^ 

£ 

/ 

o. 

/ 

/, 

/ 

/ 

\    \    \ 

/ 

/ 

OS 

,1 

f/ 

^ 

Percent  of  C 

t  of  Appa 

a 

us 

Vacuum 

100 


120    140    160 
Fia.  232. 


ISO 


200 


452 


STEAM  POWER  PLANT  ENGINEERING 


conditions  as  stated.  From  the  tables  it  would  seem  at  first  glance 
that,  except  where  coal  is  expensive,  all  the  plants  with  low  factors, 
10  per  cent  and  under,  ought  to  be  run  non-condensing.  This  is  true 
for  "  one-engine  "  installations,  but  not  necessarily  so  where  there 
are  a  number  of  engines  or  turbines.  In  the  latter  case  higher 
economy  may  be  effected  by  providing  only  a  portion  of  the  engines 
with  condensing  equipment.  The  engine  carrying  the  continuous  or 
day  load  should  operate  condensing,  and  the  non-condensing  engine 
should  carry  the  peak  load.  In  order  that  any  of  the  units  may  be 
used  for  the  day  work,  all  engines  could  be  connected  to  the  condenser, 
but  only  those  carrying  the  day  load  should  be  operated  condensing. 
Each  installation,  of  course,  must  be  considered  separately  and  due 
weight  given  to  the  various  factors  entering  into  the  problem.  For  an 
excellent  article  on  the  subject  see  " Condensers  for  Steam  Engines  and 
Turbines,"  Power,  December,  1906,  p.  769,  and  the  Engineer,  London, 
April  13,  1906,  p.  381. 

232.  Choice  of  Condensers.  —  The  proper  selection  of  a  condenser 
for  a  proposed  installation  depends  upon  the  conditions  under  which 
the  plant  is  to  be  operated.  When  there  is  a  plentiful  and  cheap  sup- 
ply of  good  condensing  water  suitable  for  boiler  feed,  and  extremely  high 
vacua  are  hot  essential,  some  type  of  jet  condenser  will  generally  be 
found  most  desirable.  If  overhead  room  permits,  a  siphon  or  baro- 
metric condenser  will  probably  be  most  suitable  and  least  expensive. 

TABLE  56. 

MOST   ECONOMICAL    VACUUM    FOR    STEAM   TURBINES. 
Vacuum  referred  to  30-Inch  Barometer. 


Load  Factor, 
per  Cent. 

Cost  of  Coal,  Dollars  per  Ton. 

$1.50 

$2.00 

$2.50 

$3.00 

$3.50 

A 

B 

A 

B 

A 

B 

A 

B 

A 

B 

5 

N.C. 

20 
24 

26.5 
27.5 
28 

N.C. 
N.C. 
17 
20 
24 
27.6 

N.C. 
23 
26.5 
27.3 

27.8 
28.2 

N.C. 
N.C. 
20 
23 
27 
27.9 

18 
25 
27 
27.6 
28 
28.3 

N.C. 
N.C. 
22 
25.5 
27.6 
28 

20 
26.5 
27.5 
27.8 
28.1 
28.4 

N.C. 
20 
24 
27 
27.8 
28 

22 
27 
27.7 
27.9 
28.2 
28.5 

N.C. 
22 
25.8 
27.5 
28 
28 

10 

15 

20  
30  

50  

A.  Surface-condensing  plant;  cost  $6  per  kilowatt  of  main  generator.    Fixed  charges  12  per  cent. 
Cost  of  water  not  included.     Rated  capacity  of  generator,  1000  kilowatts. 

B.  Surface-condensing  plant,  including  cooling  towers  and  extra  cost  of  land,  etc.;  cost  $10 
per  kilowatt  for  26-inch  plant,  increasing  to  $14  per  kilowatt  for  28.5-inch  plant.     Fixed  charges 
12  per  cent.     No  charge  for  water.     Rated  capacity  of  generator,  1000  kilowatts. 


CONDENSERS 


453 


TABLE   57. 

MOST    ECONOMICAL    VACUUM    FOR   RECIPROCATING    ENGINES. 
Vacuum  referred  to  30- Inch  Barometer 


Load  Factor, 
per  Cent. 

Cost  of  Coal,  Dollars  per  Ton. 

$1.50 

$2.00 

$2.50 

$3.00 

$3.50 

A 

B 

A 

B 

A 

B 

A 

B 

A 

B 

10 

N.C. 

16 
22.5 
24 
25.5 

N.C. 
N.C. 
N.C. 
16 
22 

15 
20 
23 

24.5 
26.7 

N.C. 
N.C. 
N.C. 
21 
23.5 

18 
22 
23.5 
25.5 
27.2 

N.C. 
N.C. 
20 
22 
23.5 

20 
22.5 
24.5 
26.4 
27.5 

N.C. 
16 
21 
23 
26.3 

22 
24 
25 

26.8 
27.7 

N.C. 
20 
22 
24 

27 

15 

20 

30 

50 

A.  Surface-condensing  plant;  cost  $7  per  kilowatt  of  main  generator.    Fixed  charge  12  per  cent. 
Cost  of  water  not  included.     Rated  capacity  of  generator,  1000  kilowatts. 

B.  Surface-condensing  plant,  including  cooling  towers  and  extra  cost  of  land,  etc.;  cost  $11  per 
kilowatt  for  26-inch  plant,  increasing  to  $13  per  kilowatt  for  27.5-inch  plant.     Other  conditions 
as  in  A. 

Where  there  is  a  plentiful  supply  of  good  water  for  boiler  feed  but 
the  water  which  must  be  used  for  cooling  purposes  is  very  dirty  the 
siphon  condenser  is  preferable  to  the  barometric  form.  A  surface  con- 
denser may  be  used  in  the  latter  case  if  the  condensing  water  is  not  so 
dirty  as  to  seriously  impair  the  efficiency  by  coating  the  tubes  with 
sediment,  and  boiler  feed  water  is  scarce. 

The  air-cooled  surface  condenser  is  employed  only  where  water  of  any 
kind  is  scarce. 

For  very  high  vacua  in  connection  with  steam  turbine  work  the  sur- 
face condenser  is  almost  universally  adopted,  although  the  barometric 
condenser  in  connection  with  dry-air  pumps  is  finding  favor  with  many 
engineers. 

In  selecting  the  type  of  condenser  and  auxiliaries  due  weight 
must  be  given  to  the  load  factor,  cost  of  coal,  water,  land,  building, 
interest,  depreciation  and  the  like,  as  outlined  in  the  preceding 
paragraph. 

233.  Water-Cooling  Systems.  —  When  an  ample  supply  of  cooling 
water  is  unobtainable  for  natural  or  economic  reasons,  the  circulating 
water  may  be  used  over  and  over  again  by  employing  suitable  cooling 
devices.  The  three  most  common  in  practice  are 

1.  The  simple  cooling  pond  or  tank. 

2.  The  spray  fountain. 

3.  The  cooling  tower. 


454  STEAM  POWER  PLANT  ENGINEERING 

233a.  Cooling  Pond.  —  The  water  is  cooled  partly  by  radiation  and 
conduction  but  principally  by  evaporation.  The  air  is  seldom  saturated 
normally,  and  its  capacity  for  absorbing  moisture  is  increased  on  account 
of  its  temperature  being  raised  by  contact  with  the  warm  water  and 
by  radiation.  The  cooling  action  is  independent  of  the  depth  of  water 
and  varies  directly  as  the  surface,  the  amount  of  heat  dissipated  for 
each  square  foot  depending  upon  the  temperature  of  the  water,  the  rela- 
tive humidity,  and  the  velocity  of  the  air  currents.  Results  of  tests  are 
very  discordant. 

Box  in  his  treatise  on  Heat  states  that  the  pond  surface  should 
approximate  210  square  feet  per  nominal  horse  power  for  an  engine 
working  twenty-four  hours  a  day.  (Treatise  on  Heat,  Box,  p.  152.) 

If  the  engine  works  only  twelve  hours  per  day,  the  area  may  be  reduced 
to  105  square  feet  per  horse  power,  because  the  water  will  cool  during 
the  night,  but  in  that  case  the  depth  should  be  such  as  to  give  a  capacity 
of  300  cubic  feet  per  horse  power.  These  figures  are  based  on  a  reduc- 
tion in  temperature  of  122  to  82  degrees  F.,  with  air  at  52  degrees  F. 
and  humidity  85  per  cent,  the  steam  consumption  per  nominal  horse 
power  being  taken  at  62.5  pounds. 

Box  gives  the  following  formula  for  the  rate  of  evaporation  in  per- 
fectly calm  air: 

E=  (243  +  3.70  (V-v),  (95) 

in  wnicn 

E  =  evaporation  in  grains  per  square  foot  per  hour. 

t  =  temperature  of  the  water,  degrees  F. 
V  =  maximum  vapor  tension  in  inches  of  mercury  at  temperature  t. 

v  =  actual  vapor  tension. 

Evaporation  is  greatly  affected  by  the  force  of  the  wind  and  varies 
from  2  to  12  times  the  amount  determined  from  equation  (95). 

Example:  How  many  pounds  of  water  will  be  evaporated  per  square 
foot  per  hour  from  a  pond  with  the  temperature  of  the  water  and  air 
80  degrees  F.;  air  perfectly  calm;  barometric  pressure  29.5  inches  and 
relative  humidity  70  per  cent? 

The  maximum  vapor  tension  at  temperature  of  80  degrees  is  1.02 
inches  of  mercury.  The  actual  vapor  tension  will  be 

1.02  X  .70  (  =  relative  humidity)  =  .714. 
Substitute  these  values  in  (95). 

E  =  (243  +  3.7  X  80)  (1.02-0.714) 
=  165  grains  per  square  foot  per  hour. 
=  .023  pound  per  square  foot  per  hour. 


CONDENSERS 


455 


If  the  temperature  of  the  water  were  130  degrees  F.  and  that  of  the 
surrounding  air  80  degrees  F.,  humidity  70  per  cent,  the  evaporation 
would  be 

E  =  (243  +  3.7  X  80)  (4.5-0.714) 
=  2040  grains  per  square  foot  per  hour 
=  0.291  pound  per  square  foot  per  hour. 

Here  4.5  =  maximum  vapor  tension,  corresponding  to  a  temperature 
of  130  degrees. 

233b.  Spray  Fountain.  —  From  equation  (95)  we  see  that  even 
under  the  most  favorable  circumstances  an  enormous  pond  surface  is 
necessary.  To  facilitate  evaporation  with  a  view  toward  reducing  the 
size  of  the  pond,  the  hot  circulating  water  is  sometimes  distributed 
through  pipes  and  discharged  through  nozzles,  falling  to  the  surface  of 
the  pond  in  a  spray.  The  following  gives  some  interesting  data  con- 


40 

J90 
,80 
}% 

I  60 

«50 
t 

«« 

JL 


ater  at  Nozzles 


10     13     1C      19     22     25 


1       4       7      10      13     1C     19      22     25 
Day-o£  the  Month    September,  1904 


1       4       7      10     13      16     19     22     25      28 
Day-of  the  Month    January,  1905 


FIG.  233.     Curves  Showing  Performance  of  Spray  Fountain;  Chattanooga  Electric 
Company's  Power  Plant. 


cerning  the  spray  fountain  installation  at  the  power  plant  of  the  Chat- 
tanooga Electric  Company,  Chattanooga,  Tenn.  (Street  Railway  Review, 
March  15,  1905.) 

Adjoining  the  power  house  a  pond  150  x  300  feet  was  excavated  to 
a  depth  of  4  feet,  the  level  of  the  water  being  8  feet  below  the  condensers . 


456  STEAM  POWER  PLANT  ENGINEERING 

Circulating  water  returned  from  the  condensers  is  distributed  through 
a  set  of  pipes  provided  with  42  nozzles  through  which  the  water  is  dis- 
charged upwards.  The  rectangle  denned  by  the  center  lines  of  the 
outermost  pipes  is  98  feet  by  125  feet.  The  pipes  are  supported  on 
brick  piers  spaced  at  intervals  of  about  20  feet  in  each  direction.  The 
discharge  opening  of  the  nozzles  is  1J  inches  in  diameter,  and  the  interior 
is  provided  with  a  spiral  core  so  that  in  its  passage  the  water  is  given 
a  rotary  motion,  the  effect  of  which  is  to  greatly  increase  the  spraying 
action.  The  nozzles,  except  on  the  extreme  outer  lines  of  piping,  are 
placed  in  pairs  with  the  axes  in  a  vertical  plane  at  right  angles  to  the 
center  line  of  the  supply  pipe,  the  axis  of  each  nozzle  making  an  angle 
of  30  degrees  with  a  vertical  plane  through  the  center  of  the  supply 
pipe.  The  effect  of  each  pair  of  nozzles  is  to  throw  a  mass  of  spray  to 
the  height  of  about  15  feet,  which  in  falling  covers  an  area  of  15  x  30 
feet. 

A  dike  extending  nearly  across  the  pond  near  one  end  provides  a 
canal  through  which  the  water  is  conducted  to  the  suction  chamber,  the 
object  being  to  draw  the  supply  from  distant  parts  of  the  pond  to  give 
greater  time  for  cooling.  The  "  make-up  "  water  is  supplied  by  wells. 
The  operation  of  the  cooling  pond  for  a  warm  month  and  for  a  cold 
month  is  shown  in  Fig.  233.  Readings  were  taken  at  three-hour  inter- 
vals. The  pond  supplies  the  circulating  water  for  three  2000-square- 
feet  Worthington  surface  condensers. 

234.  Cooling  Towers.  —  A  cooling  tower  consists  of  a  wooden  or 
sheet-iron  housing  open  at  the  top  and  bottom  and  so  arranged 
that  the  heated  cooling  water  may  be  elevated  to  the  top  and  dis- 
tributed in  such  a  manner  that  it  falls  in  thin  sheets  or  sprays  into 
a  reservoir  at  the  bottom,  air  at  the  same  time  being  drawn  in  at  the 
bottom  by  natural  draft  or  forced  in  by  a  fan.  The  water  gives  up  its 
heat  to  the  ascending  current  of  air  by  evaporation  and  conduction,  the 
latter,  however,  being  a  relatively  small  factor.  If  the  air  supply  is 
dependent  entirely  upon  convection,  the  system  is  known  as  the  natural- 
draft  or  flue  cooling  tower;  if  the  air  is  forced  into  the  tower  by  fans,  it 
is  called  a  fan  cooling  tower.  The  different  types  vary  principally  in 
the  method  of  water  distribution.  Fig.  234  illustrates  the  Barnard 
cooling  tower,  •  in  which  the  falling  water  is  broken  up  by  vertically 
suspended  galvanized  iron  wire  cloth  mats,  causing  it  to  trickle  in  thin 
sheets  to  the  bottom.  A  similar  result  is  brought  about  in  the  Worth- 
ington tower,  Fig.  235,  by  pieces  of  terra-cotta  pipe  6  inches  in  diame- 
ter and  two  feet  long  placed  on  ends  in  rows.  In  the  Alberger  cooling 
tower  the  water  trickles  down  the  sides  of  swamp-cypress  boards 
arranged  in  honeycomb  fashion.  In  the  Jennison  cooling  tower  the 


CONDENSERS 


457 


water  is  divided  into  a  rain  of  drops,  constantly  retarded  in  their  fall  by 
a  series  of  perforated  4x4  inch  galvanized  iron  trays  arranged  in 
horizontal  rows  and  staggered  vertically. 

With  the  best  forms  of  cooling  towers,  under  average  conditions,  the 


DISTRIBUTING 
TROUGH 


DISCHARGE 
FROM 
TOWER 


FIG.  234.    Barnard- Wheeler  Cooling  Tower. 


temperature  of  the  circulating  water  may  readily  be  reduced  from  40 
to  50  degrees  with  a  loss  not  exceeding  3  or  4  per  cent  of  the  total 
quantity  of  water  passing  through  the  tower.  The  power  consumed 
by  the  fan  in  a  forced-draft  apparatus  averages  2  per  cent  of  that 


458 


STEAM  POWER   PLANT  ENGINEERING 


developed  by  the  main  engines,  for  the  maximum  requirements  during 
summer  months,  and  1 J  per  cent  during  the  winter. 

The  location  of  the  tower  may  be  on  the  engine-room  floor,  on  top  of 


SS£  HOT  WATER. 


>  COLD  WATER- 


SUCTION    TANK 

FIG.  235.   Worthington  Cooling  Tower. 


the  building,  or  in  the  yard,  the  latter  being  the  most  adaptable.  It 
may  be  any  reasonable  distance  from  the  engine  and  condenser. 
Fig.  236  shows  a  typical  installation  of  Alberger  condenser  and  cooling 
towers. 


CONDENSERS 


459 


460  STEAM  POWER  PLANT  ENGINEERING 


235.  Parallel  Comparison  of  Fan  and  Natural-Draft  Cooling  Towers. 

FAN.  NATURAL  DRAFT. 

Size. 

Small,  the  forced  draft  providing  Large,  draft  being  necessarily  small, 

sufficient  air  velocity  to  effect  evapo-  a  larger  area  must  be  provided  to 

ration.  perform  same  work. 

Height  limited,  because  loss  from  back  Height  is  an  advantage  because  the 

pressure  increases  with  the  height.  tower  operates  on  the  principles  of  a 

Tower  usually  short  and  of  large  area.  chimney. 

Power  Consumption. 

One   per  cent  of  station  output  and    None. 
upwards,  depending  upon  the  type 
of  auxiliaries  and  the  conditions  of 
operation. 

Location. 

Inside  or  outside.    Can  operate  in  any  Outside  only,  unless  exceptionally  good 

location  where  sufficient  head  room  draft  is  obtainable. 

and  air  supply  are  available.  Preferably  in  the  open  where  advan- 

Especially  adapted  to  inconvenient  lo-  tage    may    be    taken    of    prevailing 

cations,  as  roofs,  upper  decks,  boiler  winds. 

floors,  etc. 

Conditions  of  Atmosphere. 

Comparatively  little  affected  by  tern-  Largely  affected  by  temperature  and 
perature,  considerably  by  humidity,  humidity  and  wind.  Draft  increased 
and  none  by  winds.  by  steady  winds. 

Conditions  of  Operation. 

More  especially  adapted  for  heavy  con-  Especially  adapted  for  light  summer 
tinuous  duty  the  year  round,  as  in  and  heavy  winter  duty,  as  in  electric- 
rail-plants  or  mills.  lighting  plants. 

First  Cost  and  Cost  of  Operation. 

First     cost     greater     on     account     of  First  cost  small  by  reason  of  simplicity 

mechanical  construction  and  neces-  and  construction. 

sary  auxiliaries.  First  cost  largely  dependent  upon  ma- 
Cost  of  operation  dependent  upon  type  terials  used  in  interior  construction. 

of  auxiliary  and  conditions  of  oper-  Cost    of    operation    limited    to    fixed 

ation.  charges. 

236.  Water-cooling    Calculations.  —  Air   is    said   to   be   completely 
saturated  when  it  contains  all  the  water  vapor  it  can  hold  without 


CONDENSERS  461 

causing  precipitation.  If  the  vapor  content  is  less  than  > that  corre- 
sponding to  complete  saturation  the  air  will  tend  to  become  saturated 
by  absorbing  moisture  from  surrounding  objects.  The  drier  the  air 
the  greater  will  be  its  affinity  for  moisture.  The  necessary  latent  heat 
for  vaporization  is  supplied  directly  by  the  water  producing  the  vapor 
or  by  the  surrounding  objects  in  contact  with  the  water.  Thus,  in  the 
open  cooling-tower  the  water  vapor  is  absorbed  from  the  circulating 
water,  and  the  heat  necessary  to  effect  this  vaporization  is  given  up  by 
the  water,  with  a  resultant  reduction  in  temperature  of  the  water  itself; 
and  in  the  evaporative  surface  condenser  the  vapor  is  absorbed  from 
the  water  spray  in  contact  with  the  tubes,  the  heat  required  to  effect 
this  vaporization  being  given  up  by  the  steam  within  the  condenser 
chambers,  resulting  in  condensation  of  the  steam.  If  the  air  coming 
in  contact  with  the  water  is  very  dry  and  at  a  high  temperature  the 
vaporization  of  the  water  may  be  rapid  enough  to  cool  the  remaining 
water  to  a  temperature  much  lower  than  that  of  the  air.  In  this  case 
practically  all  of  the  cooling  is  effected  by  evaporation.  But  when  the 
air  is  at  a  low  temperature  and  high  relative  humidity  a  considerable 
amount  of  heat  may  be  carried  away  by  the  air  by  conduction.  The 
quantity  of  air  and  water  necessary  to  produce  a  given  cooling  effect 
may  be  determined  as  follows: 

Let  H  =  total  amount  of  heat  to  be  abstracted,  B.T.U.  per  hour. 
W  =  weight  of  water  to  be  cooled,  Ibs.  per  hour. 
te  =  temperature  of  water  entering  cooling  device. 
ti  =  temperature  of  water  leaving  cooling  device. 
t0  =  temperature  of  air  entering  cooling  device,  °  F. ;  TQ  =  £0  +  460. 
t,  =  temperature  of  air  leaving  cooling  device,  °  F. ;  T^  =  t2  +  460. 
p  =  ordinary  atmospheric  pressure  =  29.92  in.  of  mercury. 
pa  =  observed  atmospheric  pressure,  in.  of  mercury. 
PQ  =  elastic  force  of  vapor  at  temperature  t0,  in.  of  mercury. 
p2  =  elastic  force  of  vapor  at  temperature  t2,  in.  of  mercury. 
F0  =  volume  of  air  entering  the  cooling  device,  cu.  ft.  per  hour, 

atmospheric  conditions. 

V2  =  volume  of  air  discharged  from  the  cooling  device  at  tem- 
perature t. 

d  =  density  of  dry  air,  at  pressure  p  and  temperature  £0. 
h0  =  weight  of  moisture  in  1  cu.  ft.  of  saturated  air  at  tempera- 
ture 4)j  pounds. 

h2  =  weight  of  moisture  in  1  cu.  ft.  of  saturated  air  at  tempera- 
ture £2,  pounds. 

20  =  relative  humidity  of  the  air  entering  the  cooling  device. 
22  =  relative  humidity  of  the  air  leaving  the  cooling  device. 


462 


STEAM  POWER  PLANT  ENGINEERING 


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CONDENSERS  463 

C  =  specific  heat  of  dry  air  at  constant  pressure,  =  0.2375. 
S  =  mean  specific  heat  of  water  vapor  at  temperature  t2. 
r2=  heat  of  vaporization  at  temperature  t2. 

The  pressure  p3  of  the  dry  air  in  atmospheric  air  entering  the  cooling 
device  is  Pi  =  Pa  -  ?Vo-  (96) 

The^pressure  p3  of  the  dry  air  leaving  the  cooling  device  is 

Pz  =  Pa-  Pfr.  (96a) 

The  weight  w  of  dry  air  entering  the  cooling  device  under  atmospheric 
temperature  and  pressure,  Ibs.  per  hour,  is 

w  =  ^dV0.  (97) 

P 

The  weight  w0  of  moisture  carried  into  the  cooling  device  by  the  air, 
Ibs.  per  hour,  is  WQ  =  h0z0V0.  (97a) 

The  volume  of  air  leaving  the  cooling  device  is 

V   —  V  —  •  — -2-  CQS^i 

>3      T0 

The  weight  w2  of  moisture  carried  away  by  the  air  discharged  from  the 
cooling  device,  Ibs.  per  hour,  is 

w2  =  h2z2V0.  (98a) 

The  weight  of  circulating  water  w3  absorbed  by  the  air  in  passing  through 
the  cooling  device,  Ibs.  per  hour,  is 

TI>S  =  w2  -  v>0.  (98b) 

The  heat  H  to  be  abstracted  from  the  circulating  water,  B.T.U.  per 
hour,  is  H  =  W(tt-te).  (99) 

The  heat  is  dissipated,  by  the  cooling  process,  in  raising  the  tempera- 
ture of  the  air  and  vapor  entering  the  cooling  device  from  t0  to  t2  (by 
conduction)  and  in  evaporating  the  moisture  absorbed  by  the  air  in  pass- 
ing through  the  apparatus  (by  evaporation). 

The  heat  Ha  required  to  raise  the  temperature  of  dry  air  from  t0  to  t2, 
B.T.U.  per  hour,  is  ffa  =  Cw  (t2-  tQ).  (100) 

The  heat  Hs  required  to  superheat  the  water  vapor  entering  the  cool- 
ing device  from  temperature  £0  to  t2,  B.T.U.  per  hour,  is 

Hs  =  w^S(t2-t0).  (lOOa) 

The  heat  Hc  abstracted  by  conduction  from  the  circulating  water, 
B.T.U.  per  hour,  is  Hc  =  Ha  +  #..  (lOOb) 

The  heat  He  abstracted  from  the  circulating  water  by  evaporation, 
B.T.U.  per  hour,  is  He  =  Wfy  (101) 


464  STEAM  POWER  PLANT  ENGINEERING 

Though  the  process  of  evaporation  is  practically  continued  through 
the  whole  range  in  the  cooling  device,  we  are  justified  in  using  the  heat 
of  vaporization  at  the  highest  temperature,  because  the  liquid  was  at 
this  temperature  entering  the  cooling  device  and  the  vapor  is  brought 
back  to  the  temperature  when  leaving  it. 

The  total  heat  Hc  absorbed  by  the  air  in  passing  through  the  cooling 
device,  B.T.U.  per  hour,  is  Ht  =  Hc  +  He.  (102) 

Neglecting  radiation  and  other  minor  losses,  the  heat  Ht  absorbed 
by  the  air  must  be  equal  to  the  heat  given  up  by  the  circulating  water,  or 

Ht  =  H.  (103) 

Example:  Determine  the  quantity  of  air  passing  through  the  cool- 
ing tower  per  hour  and  the  circulating  water  lost  by  evaporation  in  a 
power  plant  operating  under  the  following  conditions:  Engines  indicate 
500  H.P.  and  consume  20  Ibs.  steam  per  LH.P.  hour;  temperature  of 
the  injection  water,  discharge  water  and  outside  air,  90,  122  and  72°  F., 
respectively;  barometer  29.5;  relative  humidity  of  air  entering  and 
leaving  tower  70  and  90  per  cent  respectively;  vacuum  at  condenser 
25  inches.  Determine  also  the  weight  of  water  evaporated  in  per  cent 
of  that  circulated  and  of  the  condensed  steam. 
In  the  problem, 

pa  =  29.5,  to  =    72,  f  ^o  =        0.001224, 

t  Po  =    0.79,  *  t2  =  112,  f  ^2  =        0.003978, 

t  p2  =    2.79,  te  =  122,  z0  =        0.70, 

d  =    0.0747,  *,  =    90,  z2  =        0.90, 

C  =    0.2375,          S  =      0.45,  r2  =  1028.9. 

These  values  are  obtained  from  Steam  Tables  and  from  Air  Tables 
(Table  58). 
Substitute  these  values  in  equations  (96)  to  (103)  thus: 

(96),    p,  =  29.5  -  0.79  X  0.7 

=  28.95. 
(96a),p,=  29.5  -  2.74  X  0.9 

=  27.03. 
oo  qc 
(97),    w  =  —^  X  0.0747  V0. 

=  0.0722  70. 

(97a),  w  =  0.001224  X  0.7  70 
=  0.000857  V0. 

*  By  assumption,  t2  being  10  to  20  degrees  lower  than  te  in  average  practice  when 
the  range  te  —  t0  is  greater  than  30  degrees. 

t  Marks  and  Davis:  the  values  in  Table  58  are  Regnault's. 


CONDENSERS  465 

28.95460 


72     O 
=  1.152; 

that  is,  each  cu.  ft.  of  dry  air  entering  the  cooling-tower  is  increased  in 
volume  to  1.152  cu.  ft.  as  it  leaves. 

.  (98a),  w2  =  0.003978  X  0.9  X  1.152  F0 

=  0.004125  V0. 

(98b),  ws  =  0.004125  F0  -  0.000857  F0 
=  0.003268  F0. 

The  total  heat  to  be  abstracted  from  the  steam  (see  equation  (84), 

page  347)  is 

H  =  500  X  20  (1120.1  -  122  +  32) 

=  10,300,000  B.T.U.  per  hour. 
(99),  But  W  (122  -  90)  =  10,300,000, 
from  which  W  =  322,000  pounds  per  hour. 

(100),  Ha  =  0.2375  X  0.0722  F0  (112  -  72) 

=  0.6865  V0. 
(lOOa),  Hs  =  0.000857  V0  X  0.45  (112  -  72) 

=  0.001543  F0 
(100b),#c  =  0.6865  F0  +  0.001543  F0 

=  0.688  F0. 

(101),    He  =  0.003268  F0  X  1028.9 
-  3.365  F0. 


H€      3.365  F0 


4.89; 


Hc      0.688  F0 

that  is,  the  air  removes  4.89  times  more  heat  by  evaporation  than  by  con- 
duction under  the  given  conditions. 

(102),    Ht  =  0.688  F0  +  3.365  V0 

=  4.053  70. 
H.  _  0.001543  70  _ 
~Ht-      4.053  F0 

that  is,  the  heat  required  to  superheat  the  moisture  carried  into  the 
tower  by  the  air  is  approximately  j^  of  1  per  cent  of  the  total;  hence 
an  error  as  great  as  20%  in  the  mean  specific  heat  of  the  vapor  is 
negligible. 

(103),  .4.053  F0  =  10,300,000, 
from  which 

F0  =  2,543,000  cu.  ft.  of  air  per  hour 

necessary  to  effect  the  required  cooling 
=  42,300  cu.  ft.  per  minute. 


466  STEAM  POWER  PLANT  ENGINEERING 

From  (98b) 

ti>3  =  0.003268  V0. 
Substitute 

V0  =  2,543,000  in  above  equation. 

w3  =  0.003268  X  2,543,000 

=  8320  Ibs.,  or  the  weight  of  circulating  water    * 
carried  away  per  hour. 

w3  _     8320 
F~32pOO~ 

that  is,  2.58  per  cent  of  the  circulating  water  is  carried  away  by  the  air 
in  effecting  the  necessary  cooling. 

»          .   832Q      .   832. 
20  X  500       10,000 

that  is,  the  equivalent  of  83.2  per  cent  of  the  steam  used  by  the  engines 
is  evaporated  in  the  cooling  tower,  or  the  make-up  water  is  more  than 
supplied  by  the  condensed  steam. 

Example:  Evaporation  Surface  Condenser. —  How  many  cubic  feet 
of  air  and  how  many  pounds  of  water  spray  must  be  forced  through  an 
evaporative  surface  condenser  of  the  fan  type  in  order  to  condense 
1000  pounds  of  steam  per  hour  and  maintain  a  vacuum  of  25  inches, 
barometer  29?  (Atmospheric  air  80°  F.,  relative  humidity  70%.) 
The  air  and  vapor  issue  from  the  discharge  pipe  under  pressure  of  four 
inches  of  water,  temperature  120°  F.,  relative  humidity  98%. 

The  absolute  pressure  in  the  condenser  is  29.0  —  25.0  =  4  inches  of 
mercury. 

The  total  heat  to  be  withdrawn  in  order  to  cool  and  condense  1000  Ibs. 
of  steam  per  hour  at  absolute  pressure  of  4  inches  to  120°  F.  is 

1000  [1114.8  -  (120  +  32)]  =  1,026,000  B.T.U. 

Neglecting  radiation  and  leakage  losses,  this  is  the  heat  to  be  ab- 
stracted per  hour  by  the  air  and  water  spray. 

The  pressure  of  the  dry  air  in  the  mixture  entering  the  condenser  is, 

equation  (96), 

p,  =  29.0  -  0.7  X  1.029. 

=  28.28. 
The  pressure  of  dry  air  in  the  mixture  leaving  the  condenser  is, 

equation  (96a), 

p3  =  (29.0  +' 0.294)  -  0.98  X  3.438 

=  25.925 

(0.294  is  the  value  in  inches  of  mercury  of  four  inches  of  water-fan 
pressure) . 


CONDENSERS  467 


Let  F0  =  volume  of  atmospheric  air  entering  the  condenser.     The  vol- 
ume leaving  the  condenser  will  be,  equation  (98), 
,.   _  28.280    460  +  120  _ 
V°~  2^925'   460+80   = 

The  weight  of  vapor  in  the  condenser  discharge  is,  equation  (98a), 
w.2  =  1.172  F0  X  0.004888  X  0.98 

=  0.005615  F0  Ibs. 

The  weight  of  vapor  in  the  mixture  entering  the  condenser  is,  equa- 
tion (97a),  WQ  =  0.00157  X  0.7  F0 

=  0.001099  F0  Ibs. 
The  amount  evaporated  therefore  is 

w3  =  0.005615  F0  -  0.001099  F0 

=  0.004516  V0  Ibs. 
The  weight  of  dry  air  entering  the  condenser  is,  equation  (97), 


=  0.06958  F0  Ibs. 

The  heat  absorbed  by  the  dry  air  in  being  heated  from  80°  to  120°  F 
is,  equation  (100), 

H  =  Cw  (t2  -  ^ 

=  0:02375  X  0.06958  F0  (120  -  80) 
=  0.658  F0  B.T.U. 

Heat  required  to  superheat  w0  Ibs.  of  vapor  from  80°  to  120°  F.  is, 
equation  (lOOa),  ^  =  Q.001099  F0  X  0.46  (120  -  80) 

=  0.02022  F0  B.T.U. 

Heat  absorbed  by  the  evaporation  of  w3  Ibs.  of  water  is,  equation 
(101),  He  =  0.004516  F0  X  1046.7 

=  4.720  F0  B.T.U. 

(Here  the  latent  heat  is  taken  at  the  lower  temperature,  it  being  the 
original  temperature  of  the  liquid.) 

Total  heat  absorbed  by  the  entering  air  and  spray  is 
Ht  =  0.658  F0  +  4.720  F0  +  0.020  F0 

=  5.398  F0. 
But  this  represents  also  the  heat  given  up  by  the  steam,  or 

5.398  F0  =  1,026,000. 

From  which  F0  =  190,500  cu.  ft.  of  atmospheric  air  necessary  to  con- 
dense and  cool  1000  pounds  of  steam  under  the  given  conditions. 
The  water  spray  to  be  injected  per  hour  is 

0.004516  F0  =  0.004516  X  190,500  =  860  pounds. 


468  STEAM  POWER  PLANT  ENGINEERING 

236a.  Hygrometry.  —  The  degree  of  saturation,  or  relative  humidity, 
is  ordinarily  determined  from  the  difference  in  reading  of  a  wet  and  a  dry 
bulb  thermometer,  thus:  If  the  air  is  saturated  with  aqueous  vapor  no 
evaporation  takes  place  from  the  wet  bulb  and  the  two  thermometers 
give  identical  readings;  but  if  it  is  unsaturated,  evaporation  occurs. 
The  wet-bulb  thermometer  is  thus  cooled  and  its  readings  are  lower  than 
those  of  the  dry  bulb.  The  difference  in  reading  is  a  function  of  the 
relative  humidity,  and  the  latter  may  be  calculated  from  the  following 
modification  of  Apjohns'  formula: 

If  the  thermometer  reads  above  32°  F. 


If  it  reads  below  32°  F. 


, 

in  which 

h  =  relative  humidity,  per  cent. 

d  =  difference  in  reading  of  the  wet  and  dry  thermometers,  degrees  F. 
P  =  barometric  pressure,  inches  of  mercury. 

Pw  =  maximum   tension   of   aqueous  vapor  corresponding   to    the 

temperature  of  the  wet  thermometer,  inches  of   mercury. 

(This  may  be  taken  directly  from  the  Steam  Tables.) 

Pt  =  maximum  tension  of  aqueous  vapor  corresponding  to  the  tem- 

perature of  the  dry  thermometer,  inches  of  mercury. 

Example:  Determine  the  relative  humidity  when  the  dry  bulb  reads 
70°  F.,  wet  bulb  60°  F.,  barometer  28.0. 
From  the  Steam  Tables  we  find 

Pw  =  0.522;     Pt  =  0.739. 
Whence 


Tables  giving  the  relative  humidity  in  terms  of  the  temperature 
difference  are  published  in  most  engineering  handbooks  and  the  above 
calculations  are  unnecessary.  These  tables,  however,  are  based  on  a 
fixed  barometer  pressure,  whereas  the  formula  takes  the  actual  pressure 
into  consideration. 

237.  Test  of  Cooling  Tower  (Wheeler  Condenser  Company).  —  The 
following  gives  the  results  of  a  test  made  on  the  cooling-tower  plant  of 
the  A.  F.  Brown  Company  at  Elizabethport,  N.  J.  The  tower  is  work- 
ing in  connection  with  a  Wheeler  surface  condenser  of  280  square  feet 
of  cooling  surface,  mounted  over  a  10,  12X12  combined  air  and 
circulating  pumo. 


CONDENSERS  469 

Observations  made  on  June  24,  1904. 

Temperature  of  air 81  degrees 

Hygrometer 69  degrees 

Temperature  of  air  at  top  of  tower 89  degrees 

Temperature  of  water  in  troughs 105  degrees 

Temperature  of  water  in  tank 83  degrees 

Revolutions  of  fan,  239  r.p.m.,  air  pressure $  inch  water 

Velocity  of  air  out  of  tower 822  feet  per  minute 

Gallons  of  water  passing  over  mats 385  per  minute 

Vacuum 26  inches 

Temperature  of  air-pump  discharge 87  degrees 

Observations  made  June  28,  1904,  9  A.M 

Temperature  of  air .  76  degrees 

Hygrometer 59  degrees 

Temperature  of  air  at  top  of  tower 81  degrees 

Temperature  of  water  in  troughs 96  degrees 

Temperature  of  water  in  tank 78  degrees 

Revolutions  of  fan,  232  r.p.m.,  air  pressure §  inch  water 

Velocity  of  air  out  of  tower 680  feet  per  minute 

Gallons  of  water  passing  over  mats 406  per  minute 

Vacuum 25.5  inches 

Temperature  of  air-pump  discharge 90  degrees 

Observations  made  June  28,  1904,  3  P.M. 

Temperature  of  air 74  degrees 

Hygrometer 57  degrees 

Temperature  of  air  at  top  of  tower 83  degrees 

Temperature  of  water  in  troughs 99  degrees 

Temperature  of  water  in  tank 80  degrees 

Revolutions  of  fan,  237  r.p.m  ,  air  pressure ^  inch  water 

Velocity  of  air  out  of  tower 769  feet  per  minute 

Gallons  of  water  passing  over  mats 470  per  minute 

Vacuum 25.5  inches 

Temperature  of  air-pump  discharge 92  degrees 

Observations  made  June  29,  1904. 

Temperature  of  air 78  degrees 

Hygrometer 71  degrees 

Temperature  of  air  at  top  of  tower 86  degrees 

Temperature  of  water  in  troughs 108  degrees 

Temperature  of  water  in  tank 82  degrees 

Revolutions  of  fan,  241  r.p.m.,  air  pressure f  inch 

Velocity  of  air  out  of  tower 772  feet  per  minute 

Gallons  of  water  passing  over  mats .. 430  per  minute 

Vacuum       25.5  inches 

Temperature  of  air-pump  discharge 93  degrees 

Specifications  for  condensers   —  See  paragraph  414. 

RESULTS  OF  TEST  OF   NATURAL-DRAFT  TOWER,  DETROIT. 
COMPLETE   FIVE-FIFTHS  SURFACE  INSTALLED. 

Proc.  A.S.M.E.,  Mid-Nov.,  1909,  p,  1205. 
Engines:  Two  400-i.h.p.  300-kw.  Macintosh  &  Seymour  tandem-compound 

engines,  overhung  generators. 

Condensers:       Worthington  surface  (admiralty  type)  1600-sq.  ft.  reciprocating  wet- 
air  pump  and  circulating  pump 
Tower:  Wood-mat  construction,  24,500  sq.  ft  evaporating  surface,  exclusive 

of  shell. 
Test:  March  15  to  16,  1901,  4  p.m.  to  4  p  m  ,  24  hr. 


470 


STEAM   POWER   PLANT   ENGINEERING 


A.M.                                 P.M.  AVERAGE. 

Weather:  Barometer  (abs.),  min 30.22         30.07;     30.14         30.27 

Temperature  air,  deg 18.5           25;            30  25 

Relative  humidity,  per  cent  .        76               82;           58  72 

Load:                  600  kw.  max.  to  50  kw.  min.     Average 244.9  kw. 

Engine  efficiency  =  92.5  =  875  i.h.p.  max.      Average  .  .354.8  i.h.p. 

Steam:                Weight  of  condensed  steam  per  hr.,  Ib 5910 . 6 

Temperature  exhaust  steam,  deg.  F 134 . 38 

Temperature  condensed  steam,  deg.  F 108. 78 

Weight  of  steam  per  hour,  max.  load,  Ib 13,500 

Vacuum  (abs.)  25  to  19,  average  about 22 

Vacuum  corresponding  to  temperature  exhaust  steam. .  .  25 

Vacuum  possible  with  good  condenser  (10  dog.  difference)  28 

Water:                Circulated  per  hr.,  Ib 293,536 

Temperature  hot  well,  average,  deg.  F 87.50 

Temperature  cold  well,  average,  deg.  F 71.27 

Vaporization  loss  per  hr.,  Ib 5970 

Results:              Condenser  surface  per  kw.,  sq.  ft 2.66 

Steam  per  kw.  hr.,  Ib 24.3 

Steam  per  i.h.p.  hr.,  Ib 16.66 

Circulating  water  per  Ib.  of  steam,  Ib ". '.'.''  49.6 

Steam  per  sq.  ft.  condenser  surface  per  hr.,  Ib 3.7 

Circulating  water  per  sq.  ft.  tower  surface,  Ib 12 

Difference  in  temperature  between  exhaust  steam    and 

discharge,  deg.  F 47 

Cooling:              Max.  20  deg.,  min.  3  deg.-5  deg.     Average 16.23 

Heat  dissipated  per  hr.,  B.T.U 4,769,000 

Heat  per  sq.  ft.  tower  surface,  B.T.U 195 

Heat  per  sq.  ft.  per  1000  Ib.  water,  B.T.U 0.665 

Evaporation :     Circulating  water,  per  cent 2 .03 

Engine  steam,  per  cent 101 

Tower:                Surface  per  kw.  (average  load  245  kw.),  sq.  ft 100 

Surface  per  kw.  (max.  load  600  kw.),  sq.  ft 40.8 

Surface  per  1000  Ib.  steam  max.  load,  sq.  ft 1820 

Surface  per  1000  Ib.  steam  average  load,  sq.  ft 4140 

Surface  per  1000  Ib.  circulating  water  per  deg.  max.  cool- 
ing, sq .  ft 4.17 


Time. 


Temperature,  Deg.  Fahr. 


Air. 


Hot 

Well.* 


Cold 
Well. 


Water 
Cool- 
ing. 


Total 

Heat 

Head.t 


Quantities. 


Tower 

Water,  Lb. 

per  Hr. 


Heat  Dissi- 
pated, B.t.u. 
Lb.  per  Hr. 


8 


10 


Load, 
Kw. 


11 

270 

)315 

1290 

315 

350 

365 
485 
655 
570 
600 


12 noon 
1.30 

2.30 
3.30 

4.30 
5.00 
6.00 
7.00 
8.00 


34 
35 

35 
35 

32.5 
28.5 
26 

24 
24 


102 
106.5 

106.5 
113 

100 

103.5 

125 

121 

123 


89 
90 

87.5 
88.5 

84 
88 
94 
94 
94.5 


13 
16.5 

19 
24.5 

16 

15.5 
31 
27 

28.5 


71.5 

71.5 

78 

67.5 

75 

99 

97 
99 


375,000 

+  (375,000 

(370,200 

375,000 

375,000 

399,000 
445,500 
417,000 
427,000 
427,000 


4,880,000 
6,108,000 

7,120,000 
9,000,000 

6,384,000 

6,900,000 

12,930,000 

11,532,000 

12,174,000 


332 
415 

484 
613 

434 
470 
880 

785 
827 


25 
24.8 

25 
25 

26.6 
29.7 
27.8 
27.4 
27.4 


*  Assuming  a  more  efficient  condenser,  say  10  deg.  difference,  the  probable  vacuum  would  be 
26  deg.  to  27.5  deg.     This  condenser  actually  operated  at  40  deg.  to  50  deg.  difference. 

t  Total  heat  head  =  air  heating  4-  lost  head.  J  Difference  due  to  rapid  change  in  load. 


CHAPTER  XII. 

FEED-WATER  PURIFIERS  AND  HEATERS. 

238.  General.  —  All  natural  waters  contain  more  or  less  foreign 
matter  either  in  suspension  or  solution.  Waters  containing  carbonates 
and  sulphates  of  magnesia  and  lime,  soluble  salts  of  silica,  iron,  and 
alumina,  and  suspended  matter,  tend  to  form  scale  in  the  boiler 
and  reduce  its  steam-generating  capacity  and  economy.  The  loss  due 
to  this  cause  is  often  overestimated  but  is  of  secondary  importance 
to  the  danger  due  to  retarded  heat  transmission  which  overheats  and 
weakens  the  plates  and  tubes. 

Table  59  gives  the  results  of  a  number  of  tests  made  on  loco- 
motive boiler  tubes  with  different  thicknesses  and  characters  of 
scale.  The  diversity  of  the  results  indicates  the  futility  of  bas- 
ing the  decrease  in  conductivity  on  the  thickness  of  the  scale.  For 
example,  test  No.  1  shows  a  decrease  in  conductivity  of  9.1  per 
cent  for  a  scale  .02  inch  thick,  while  No.  16  shows  a  decrease 
of  only  6.75  per  cent  for  a  scale  over  6.5  times  as  thick.  The 
scale  in  each  case  was  even,  hard,  and  dense.  Again,  No.  8  with 
a  very  soft  scale  .042  inch  thick  gives  a  decrease  in  conductivity 
of  9.54  per  cent,  whereas  No.  14,  also  very  soft  but  twice  as  thick,  gives 
a  decrease  of  only  4.95  per  cent.  No  doubt  the  heat  transmission  is  a 
function  of  the  chemical  as  well  as  the  physical  properties,  but 
further  experiments  are  necessary  before  any  specific  conclusion  can  be 
drawn. 

Waters  containing  acids,  organic  matter,  and  magnesium  chloride 
and  sulphate  tend  to  corrode  the  boiler,  and  those  containing  sodium 
carbonate,  organic  matter,  and  alkalies  induce  priming.  Even  distille4 
water,  as  obtained  from  a  surface  condenser,  is  a  solvent  of  iron  to  a 
certain  extent  and  causes  corrosion  and  pitting.  Table  60  gives  some 
idea  of  the  character  and  extent  of  impurities  in  water  from  various 
localities,  with  an  analysis  of  the  scale  produced  by  the  water  and  the 
trouble  in  the  boiler  arising  from  its  use. 

It  is  impossible  to  judge  the  quality  of  feed  water  merely  by  the 
grains  of  solids  per  gallon,  since  a  large  amount  of  soluble  salt  such  as 
sodium  chloride  will  not  be  as  deleterious  as  a  very  small  amount  of 
calcium  sulphate. 

471 


172 


STEAM  POWER  PLANT  ENGINEERING 


TABLE   59. 
INFLUENCE    OF    SCALE    ON    HEAT    TRANSMISSION. 

(Locomotive  Boiler  Tubes.) 


No. 

Thickness  of  Scale, 
Inches. 

Character  of  Scale. 

Decrease  in  Con- 
ductivity due  to 
Scale.     Per  cent. 

1  

02 

Hard   dense 

9   1 

2  

02 

Hard 

2  02 

3 

033 

Soft 

4  3 

4 

033 

Very  hard 

3  5 

5   . 

038 

Medium 

4  03 

6   

04 

Soft   porous 

6  82 

7  

04 

Hard   dense 

3  07 

8  

042 

Very  soft 

9  54 

o,  

047 

Hard 

2  75 

10  

065 

Medium 

2  39 

11  

07 

Soft 

2  38 

12 

07 

Hard 

4  43 

13 

085 

Soft   porous 

19  0 

14 

089 

Very  soft 

4  95 

15 

11 

Hard   porous 

16  73 

16 

13 

Hard   dense 

6  75 

From  tests  conducted  at  the  University  of  Illinois,  Railroad  Gazette,  Jan.  27,  1899,  June  14, 
1901.  See  also  Engineering  Record,  Jan.  14,  1905,  p.  53;  Power,  February,  1903,  p.  70; 
Street  Railway  Review,  July  15,  1901,  p.  415. 

The  following  is  a  rough  rating  according  to  the  number  of  grains  of 
incrusting  solids  per  United  States  gallon: 

Less  than 

8  grains very  good. 

12  to  15  grains good. 

15  to  20  grains fair. 

20  to  30  grains bad. 

Over  30  grains , very  bad. 

This  applies  to  calcium  carbonate,  magnesium  carbonate,  and  mag- 
nesium chloride.  For  water  containing  sulphate  of  calcium  and  mag- 
nesium, divide  the  first  column  by  4  for  the  same  rating. 

On  account  of  the  great  variety  of  possible  impurities  the  proper 
treatment  to  be  adopted  can  be  determined  only  by  chemical  analysis 
of  the  feed  water  in  each  case. 

Table  61,  compiled  by  the  Hartford  Steam  Boiler  Inspection  and 
Insurance  Company,  shows  the  number  of  boilers  inspected  by  that 
company  during  the  year  1907  and  the  number  found  defective  from 
various  causes. 


FEED-WATER  PURIFIERS  AND  HEATERS 


473 


K 


M 


Lake 
Michigan 
Chicago. 


Water  Analysis.  Grai 
Gallon. 


.THOTH 

t"1  1—  1 


OiOOOOOOOiOOi-HCN 

cot^io-'tf-'^THCNoco 

COOrHlOOOCOOt^O 


O5  CO  Oi  O  O3  -,  O 
COT-HIOO5  ^Tt< 
t^i—  IIOO51O  2I>- 


.r-i 

l-H 


r-l  0  T-l  T-l  0  0  T-H 


T-H 


OS'-HIMCM....  OO 

oseosooi^f^co 
Ot-'-CiOis'OO 

03       .. 


of  iron  and  al 
ate  of  lime 
te  of  lime 
a  of  magnes 
d  potassiu 
d  potassiu 
atter 
al  matte 
nesia 


a 
m 


te 
an 
an 
m 
n 
o 


Silica 
Oxide 
Carbon 
Sulpha 
Carbon 
Sodiu 
Sodium 
Organic 
Total  m 
Chloride 


_, 

H     OQ 
S    3 


CD'S         iO<NCJO5rs-'eOt>i 


O-^CX) 
CXDt^CO 


O5  CO  O5  »O  rl  T-I  CO       •  TH  CO 


i  SB    .    . 

'H^S 


J-         (»  ^  «O    ^  O 
^3         -^^OS    §» 

."S       ^H  d  o  ^  <M' 


^- 


eg  ^ 

Wtn 


• 
fl  03 


o    -  a 

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-  ' 


1303 


^-i    °2    C 


w 

3  ft 


pip 

«S«4_.  "tr  >  he  , 


1PI 

SS'rt  fe  a: 


-S    "5 


« 


474 


STEAM  POWER  PLANT  ENGINEERING 


TABLE  61. 

SUMMARY    OF    INSPECTORS'   REPORTS    FOR   THE    YEAR    1907. 
(Hartford  Steam  Boiler  Inspection  and  Insurance  Company.) 


Nature  of  Defects. 


Whole  Number. 


Dangerous. 


Cases  of  deposit  of  sediment 18,917 

Cases  of  incrustation  and  scale 38,427 

Cases  of  internal  grooving 3,010 

Cases  of  internal  corrosion 12,802 

Cases  of  external  corrosion 10,230 

Defective  braces  and  stays 2,219 

Settings  defective 6,363 

Furnaces  out  of  shape 7,564 

Fractured  plates 3,551 

Burned  plates 4,878 

Laminated  plates   898 

Cases  of  defective  riveting 3,582 

Defective  heads , 1,764 

Leakage  around  tubes 1 1,357 

Cases  of  defective  tubes 8,266 

Tubes  too  light 1,947 

Leakage  at  joints 5,557 

Water  gauges  defective 3,008 

Blow-offs  defective 4,216 

Cases  of  deficiency  of  water 413 

Safety  valves  overloaded 1,231 

Safety  valves  defective 1,211 

Pressure  gauges  defective 7,651 

Without  pressure  gauges 194 

Unclassified  defects 27 

Totals.  .  159,283 


1,315 

1,333 

258 

528 

768 

578 

699 

396 

568 

499 

92 

823 

238 

1,599 

3,054 

563 

430 

707 

1,250 

156 

415 

407 

465 

194 

10 

17,345 


The  neutralization  or  elimination  of  the  impurities  may  be  effected 
by  one  of  the  following  methods: 

1.  Chemically. 

Boiler  compounds. 
Purifying  plants. 

2.  Mechanically. 

Filters. 
Blow-off. 
Tube  cleaners. 

3.  Thermally. 

Feed-water  heater. 
Distillation. 


FEED-WATER  PURIFIERS  AND  HEATERS 


475 


The  following  chart  ("  Boiler  Waters,"  W.  W.  Christie)  outlines  some 
of  the  troubles  arising  from  feed  water,  their  cause  and  means  for 
preventing  them. 


Trouble. 


Cause. 


Remedy  or  Palliation. 


Incrustation. 


Corrosion. . 


Priming. 


Sediment,  mud,  clay,  etc..  .  | 
Readily  soluble  salts 

Bicarbonate  of  magnesia,       I 
lime,  iron j 

Organic  matter 

Sulphate  of  lime j 

Organic  matter < 


Grease j 

Chloride    or    sulphate    of     I 

magnesium \ 

Sugar | 

Acid f 

Dissolved  carbonic  acid  and  S 
oxygen ^ 

Electrolytic  action 

Sewage -j 

Alkalies 

Carbonate  of  soda  in  large  ) 
quantities } 


Filtration. 

Blowing  off. 

Blowing  off. 

Heating  feed  and  precipitate. 

austic  soda. 
Lime. 
Magnesia. 
See  below. 
Sodium  carbonate. 
Barium  chloride. 
Precipitate  with  alum  i 
Precipitate  withferric  J  and  filter 
chloride  ) 

^rbonatTof  soda  }  and  filter 
Carbonate  of  soda. 

Alkali. 

Slaked  lime. 

Caustic  soda. 

Heating. 

Zinc  plates. 

Precipitation  with  alum  or  ferric 

chloride  and  filter. 
Heating  feed  and  precipitate. 

Barium  chloride. 


Analysis  of  Water  for  Softening  by  Chemical  Processes:  Jour.  Soc.  Chem.  Ind., 
May,  1899,  p.  520.  Volumetric  Determination  of  Calcium  and  Magnesium  in  Water: 
Jour.  Soc.  Chem.  Ind.,  May  31,  1901,  p.  507.  Simple  Tests  for  Boiler  Water:  Soc. 
Engng.,  May,  1904,  p.  238.  Qualitative  Tests  of  Feed  Water:  Power,  Dec.,  1904, 
p.  756.  New  Testing  Apparatus  for  Boiler  Feed  Water:  West.  Elecn.,  Aug.  1,  1903, 
p.  85.  A  Simple  Method  of  Calculating  Water  Analyses  and  Amounts  of  Substances 
to  be  Added  for  Preventing  Scale  and  Corrosion:  Jour.  Frank.  Inst.,  Vol.  CLIX,  p.  217. 
Description  of  Dearborn  Drug  and  Chemical  Co.'s  Laboratories  for  Analyses  of  Boiler 
Feed  Water:  St.  Ry.  Review,  Sept.  15,  1901. 

Boiler  Corrosion:  Power,  Jan.,  1906,  Oct.,  1905,  p.  591,  Dec.,  1900;  Eng.  Rev., 
Oct.,  1904,  p.  12;  Eng.  Mag.,  Dec.,  1905,  p.  425,  Oct.,  1900,  p.  118;  Engng.,  Oct.  10, 
1902,  p.  482;  Engr.  U.S.,  Jan.  1,  1907,  p.  103,  May  1,  1902,  p.  388;  Engr.,  Lond., 
Aug.  5,  1904,  p.  131,  July  29,  1904,  p.  115,  Dec.  4,  1896,  p.  574;  Elec.  Age,  Dec., 
1905,  p.  456;  Am.  Elecn.,  Aug.,  1905,  p.  436,  April,  1902,  p.  184;  Jour.  Am.  Soc. 
Nav.  Engrs.,  May,  1902;  Mines  and  Min.,  Sept.,  1903;  Stahl  u.  Eisen,  Jan.  15,  1904. 

Boiler  Incrustation:  Am.  Elecn.,  April,  1904,  p.  206,  Dec.,  1901,  p.  576,  May, 
1901,  p.  220,  Oct.,  1898,  p.  473;  Cassier's  Mag.,  July,  1903,  p.  273;  Chem.  News, 


476  STEAM  POWER  PLANT  ENGINEERING 

Oct.  18,  1901,  p.  191;  Engr.,  Lond.,  Jan.  21,  1898,  p.  52;  Engr.  U.S.,  April  16,  1906, 
March  15,  1904,  p.  202,  May  15,  1904,  p.  354,  Sept.  1,  1904,  p.  608;  Ice  and  Refrig., 
Nov.,  1905,  p.  173.  Chemistry  of  Scale:  Jour.  Frank.  Inst.,  Aug.,  1901,  p.  113,  Aug. 
1891,  p.  145.  Boiler  Scale:  Power,  Dec.,  1905,  p.  779 ;  St.  Ry.  Review,  April  2,  1904, 
p.  545.  Scale  Prevention:  Am.  Elecn.,  Dec.,  1901,  p.  578;  Am.  Engr.,  May,  1900, 
p.  138;  Eng.  Mag.,  1897,  12-959,  13-74,  232,  419;  Elec.  Engr.,  Lond.,  July  20,  1900, 
p.  91;  Engrs.  Gaz.,  July,  1902. 

Foaming.  —  Foaming  Water  and  Scaling  Water  for  Locomotive  Boilers:  Eng. 
News,  July  21,  1904,  p.  71,  Sept.  1,  1904,  p.  198;  Foaming  and  Priming:  Soc.  Nav. 
Arch,  and  Marine  Engrs.,  1902;  R.R.  Gazette,  Oct.  12,  1900;  Christie,  Boiler  Waters, 
Chap.  V;  Stromeyer,  Steam  Boilers,  pp.  67-83;  Rowan,  Modern  Steam  Boilers, 
pp.  321-354. 

239.  Chemical    Purification.  —  Chemical    treatment    of    boiler    feed 
water   has    been  remarkably    developed   during   recent    years    and   a 
number  of  manufacturing  concerns  make  this  their  sole  business.     The 
two  most  common  systems  of  chemical  treatment  involve  (1)  boiler 
compounds  and   (2)  purifying   plants.     In   the   former   the  necessary 
chemical  action  takes  place  inside  the  boiler  and  in  the  latter  the 
water   is  purified   before   it    enters    the   boiler.     In   either    case   the 
usual  procedure  is  to  submit  for  analysis  a  sample  of  the  feed  water 
and  the  resulting  scale  to  a  competent  chemist  who  will  specify  the 
character   and  quantity  of   chemicals   necessary  to   bring   about   the 
desired  result. 

240.  Boiler  Compound.  —  The  object  of  treatment  with  boiler  com- 
pounds is  to  neutralize  the  evil  effects  of  the  impurities  in  the  feed 
water  or  to  change  them  into  others  which  are  less  objectionable  and 
which   are   easily  removed.     When  properly  compounded   and  intro- 
duced into  the  boiler  such  preparations  are  of  great  benefit  and  prac- 
tically overcome  the  deleterious  effects,  but  when   improperly    used 
they  may  produce  even  greater  troubles  than  the  impurities  which  they 
are  expected  to  eliminate. 

Boiler  compounds  may  be  divided  into  three  classes: 

1.  Those  converting  the  scale-forming  elements  into  new  sub- 
stances which  will  not  form  a  hard,  resisting  scale  and  which  are 
readily  removed  by  skimming,  blowing  off,  or  by  tube  cleaners.  For 
example,  feed  water  containing  sulphates  of  lime  and  magnesia  will 
form  a  dense,  tenacious  scale.  If  carbonate  of  soda  be  added  in  correct 
amount,  the  sulphates  are  converted  into  insoluble  carbonates  which 
are  precipitated  and  form  scale  varying  from  a  more  or  less  porous, 
friable  crust  to  a  soft  "  mush  "  or  mud.  The  resulting  sulphate  of 
soda  remains  in  solution  and  does  not  form  scale  unless  allowed  to 
concentrate,  and  this  is  prevented  by  blowing  off.  An  excess  of  soda 


FEED-WATER  PURIFIERS  AND   HEATERS  477 

is  apt  to  cause  foaming  and  at  high  temperatures  is  liable  to  attack 
the  inside  of  gauge  glasses.  Bisodium  and  trisodium  phosphate, 
sodium  tannate,  fluoride  of  sodium,  sugar,  etc.,  have  all  proved  satis- 
factory, but  as  each  case  requires  special  treatment  no  detailed  dis- 
cussion is  possible  within  the  scope  of  this  work  and  the  reader  is 
referred  to  the  accompanying  bibliography. 

2.  Those   enveloping  the   newly  precipitated  scale-forming  crystals 
with  a  surface  which  prevents  them  from  cementing  together.     The 
ingpedients  used  to  bring  about  this  result  are  starches,  woody  fibers, 
dextrine,  slippery  elm,  and  the  like. 

3.  Those  preventing  the   formation  of  hard  scale  by  a  solvent  or 
"  rotting  "  action,  as  kerosene  and  petroleum  oils. 

Boiler  Compounds.  —  Use  of  Compounds:  Eng.  News,  July  27,  1905,  p.  112; 
Am.  Mach.,  Dec.  7,  1899,  p.  115,  Oct.  26,  1899,  p.  1014;  Power,  Aug.,  1903;  Eng. 
and  Min.  Jour.,  Aug.  12,  1905,  p.  253. 

241.  Use  of  Kerosene  and  Petroleum  Oils  in  Boiler  Feed  Water.  — 

Kerosene  oil  and  other  refined  petroleum  oils  are  sometimes  used  with 
good  effect  in  boilers  to  prevent  scale  from  adhering.  These  oils  are 
said  to  change  the  deposit  of  lime  from  a  hard  scale  to  a  friable  material 
which  may  be  easily  removed.  They  are  ordinarily  fed  to  the  boiler 
with  the  feed  water,  drop  by  drop,  through  a  sight  feed  apparatus 
similar  to  a  cylinder  oil  lubricator.  From  extended  experiments 
made  on  a  100-horse-power  tubular  boiler  fed  with  water  containing  6.5 
grains  of  solid  matter  per  gallon  it  was  found  that  one  quart  of  kero- 
sene per  day  was  sufficient  to  keep  the  boiler  entirely  free  from  scale. 
Prior  to  the  introduction  of  the  oil  the  water  had  a  corrosive  action 
upon  some  of  the  fittings  attached  to  the  boiler,  but  after  the  oil  had 
been  used  for  a  few  months  it  was  found  that  the  corrosive  action  had 
ceased.  In  another  case  40  gallons  of  kerosene  were  used  in  24  hours 
in  a  steamer  of  about  3000  horse  power.  These  boilers  showed  no 
incrustation  but  considerable  corrosion.  Evidently  oil  does  not  have 
the  same  effect  or  give  the  desired  results  in  all  cases.  Kerosene  used 
in  moderate  quantities  will  not  cause  foaming.  Crude  oil  should  never 
be  used,  as  the  heavy  residue  causes  the  formation  of  a  tough,  imper- 
vious scale  productive  of  bagged  sheets  and  collapsed  flues. 

Use  of  Kerosene  in  Boilers:  Engr.  U.S.,  Sept.  15,  1905,  p.  634;  Eng.  News,  May 
24,  1890,  p.  497;  Power,  Aug.,  1895,  p.  13,  May,  1896,  p.  16^  Trans.  A.S.M.E.,  9-247, 
11-937;  Locomotive,  July,  1890,  p.  97. 

242.  Use  of  Zinc  in  Boilers.  —  Zinc  is  often  introduced  into   boilers 
to  prevent  corrosion.     The  theory  is  that  a  feeble  but  continuous  cur- 


478  STEAM  POWER  PLANT  ENGINEERING 

rent  of  hydrogen  is  generated  over  the  whole  extent  of  the  iron  by 
electrolytic  action.  The  bubbles  of  hydrogen  formed  isolate  the 
metallic  surface  from  scale-forming  substances.  If  there  is  but  a  little 
of  the  scale-forming  element  it  is  precipitated  and  reduced  to  mud;  if 
there  is  considerable,  coherent  scale  is  produced  which  takes  the  form 
of  the  iron  surface  but  does  not  adhere  to  it,  being  prevented  from  doing 
so  by  the  intervening  bubbles  of  hydrogen.  Zinc  is  ordinarily  sus- 
pended in  the  water  space  of  the  boiler  in  the  shape  of  blocks,  slabs, 
or  as  shavings  in  a  perforated  vessel.  Electrical  connection  between 
the  metallic  surfaces  is  essential.  Rolled  zinc  slabs  12x6xJ  inches 
have  found  much  favor  in  marine  practice.  Generally  speaking  one 
square  inch  of  zinc  surface  is  sufficient  for  every  50  pounds  of  water 
in  the  boiler,  though  the  quantity  placed  in  the  boiler  should  vary  with 
the  hardness.  The  British  Admiralty  recommends  the  renewing  of  the 
zinc  slabs  whenever  the  decay  has  penetrated  to  a  depth  of  J  inch 
below  the  surface.  Zinc  does  not  prevent  corrosion  or  scale  formation 
in  all  cases  and  may  even  aggravate  the  trouble. 

Use  of  Zinc  in  Boilers:  Am.  Elecn.,  Dec.,  1901,  p.  572;  Kent,  Steam  Boiler  Econ- 
omy, p.  318;  Christie,  Boiler  Waters,  p.  137;  Stromeyer,  Marine  Boiler  Management 
and  Construction,  p.  81. 

243.  Methods  of  Introducing  Compounds.  —  Boiler  compounds  may 
be  introduced  into  the  boiler  continuously  or  intermittently.     Small 
quantities   introduced   continuously   or   at   short   intervals   are   more 
effective  than  large  quantities  at  long  intervals.     Continuous  feeding 
is  ordinarily  brought  about  by  connecting  the  suction  side  of  the  feed 
pump  with  a  reservoir  containing  the  compound  in  solution,  arranged 
similarly  to  an  ordinary   cylinder  oil  lubricator.     In  large   plants  an 
independent  pump  is  often  used  to  force  the  solution  into  the  feed  line. 
Intermittent  feeding  is  brought  about  by  temporarily  connecting  the 
suction  of  the  feed  pump  with  the  reservoir  containing  the  compound. 
The  use  of  boiler  compounds  does  not  necessarily  prevent  scale  from 
forming  in  time,  though  it  will  reduce  the  evil  to  a  minimum.     In  some 
instances  where  compounds  are  used  it  is  found  necessary  to  run  a 
tube  cleaner  through  the  tubes  at  certain  intervals,  in  others  such  a 
course  has  not  been  found  necessary. 

244.  Weight   of   Compound   Required.  —  The   weight   of   compound 
introduced  depends  upon  the  nature    of  the  reagents  used  and  the 
character  of  the  feed  water,  and  ranges  from  a  few  ounces  to  several 
pounds  per  100  gallons  of  feed  water.     For  example,  water  containing 
4  grains  of  calcium  sulphate  and  6  grains  of  magnesium  sulphate  per 
gallon,  will  require  3.57  pounds  of  carbonate  of  soda  per  1000  gallons 


FEED-WATER  PURIFIERS  AND  HEATERS  479 

of  water  for  the  reduction  of  the  sulphates.     The  chemical  reaction 
and  analysis  is  as  follows: 

CaSO4  +  Na2CO3  =  CaCO3   +  Na2SO4 
MgS04  +  Na2CO3  =  MgCO3  +  Na2SO4 

If  x  =  grains  of  Na2CO3  necessary  for  the  calcium, 

CaS04  :  Na2CO3  +  10H2O  =  4  :  x. 

40  +  32  +  4  X  16  :  2  X  23  +  12  +  3  X  16  +  10  (2  +  16)  -  4  :  x. 

x  =  8.41  grains. 

If  y  =  grains  of  Na2CO3  necessary  for  the  magnesium, 

MgS04  :  Na2C03  +  10H20  =  6  :  y. 

24  +  32  +  4  X  16  :  2  X  23  +  12  +  3  X  16  +  10  (2  +  16)  =  6  :  y. 

y  =  14.3. 

The  total  weight  of  carbonate  of  soda  per  1000  gallons  is  therefore 

1000  (14.3  +  8.41)  =  22,710  grains. 
=  3.24  pounds. 

This  amount  would  effect  the  desired  result  if  the  chemical  reaction 
is  permitted  to  take  place  for  some  time,  otherwise  an  excess  of  reagent 
is  necessary. 

245.  Mechanical     Purification.  —  Waters     containing    sand,     mud, 
organic  matter,  and  in  fact  all  matter  which  is  not  in  solution  or  in 
chemical  combination  with  the  water  may  be  purified  by  mechanical 
filtration.     Mud  and  sand  may  be  eliminated  by  simply  permitting  the 
water  to  stand  for  some  time  in  settling  tanks.     Suspended  matter 
which  will  not  gravitate  to  the  bottom  may  be  removed  by  filtering  the 
water  through  coke,  cloth,  excelsior,  or  the  like.     Filters  should  be  in 
duplicate  for  continuity  of  operation. 

Vegetable  and  other  organic  impurities  commonly  float  on  the  sur- 
face of  the  water  when  the  boiler  is  making  steam,  and  may  be  blown 
out  through  a  "  surface  blow-out."  (See  paragraph  82.) 

Precipitated  matter  may  be  ejected  from  the  boiler  by  fre- 
quent blowing  off  before  it  has  time  to  adhere  and  bake  to  a  crust. 
This  procedure  is  particularly  essential  when  boiler  compounds  are 
used. 

For  description  and  use  of  mechanically  operated  tube  cleaner  see 
paragraph  86. 

246.  Thermal  Purification.  —  (See  also  Live  Steam  Purifiers,  para- 
graph 263.)     The  carbonates  of  lime  and  magnesia  are  held  in  solution 


480  STEAM  POWER  PLANT  ENGINEERING 

in  fresh  water  by  an  excess  of  carbon  dioxide  and  are  completely  pre- 
cipitated by  boiling.  At  ordinary  temperatures  carbonate  of  lime  is 
soluble  in  approximately  20,000  times  its  volume  of  water,  at  212 
degrees  F.  it  is  slightly  soluble,  and  at  290  degrees  it  is  insoluble.  Sul- 
phate of  lime  is  much  more  soluble  in  cold  than  in  hot  water,  and  is  com- 
pletely precipitated  at  290  degrees.  (Revue  de  Mecanique,  November, 
1901,  pp.  508,  743.)  Thus  it  will  be  seen  that  a  feed  heater  may  be 
relied  upon  to  remove  part  or  all  of  the  lime,  depending  upon  the  tem- 
perature to  which  the  water  is  raised  and  the  time  in  which  the  pre- 
cipitation is  permitted  to  take  place. 

Influence  of  Temperature  and  Concentration  on  the  Saline  Constituents  of  Boiler 
Water :  Jour.  Soc.  Chem.  Ind.,  Oct.  31,  1900,  p.  885.  Solubility  of  Sulphate  of  Lime  : 
Rev.  de  Mecanique,  Jan.,  1901,  p.  5,  Nov.,  1901,  p.  508. 

247.  Purifying  Plants.  —  The  function  of  a  purifying  plant  is  the 
elimination  of  all  impurities  from  the  feed  water  before  it  enters  the 
boiler.  In  the  Scaife  system  for  water  purification  feed  water  first 
enters  the  heater,  where  it  attains  a  temperature  of  from  200  to  210 
degrees  F.  As  a  portion  of  the  free  CO2  is  driven  off  by  the  heat  the 
carbonates  of  lime  and  magnesia  are  precipitated  and  are  deposited  in 
removable  pans  inside  the  heater.  On  its  way  the  heated  water  is 
forced  by  the  boiler  feed  pump  into  a  large  precipitating  tank,  where 
the  necessary  chemicals  are  introduced  by  means  of  two  small  pumps. 
These  pumps  take  the  solution  of  chemicals  from  the  solution  tanks  which 
hold  a  sufficient  quantity  to  operate  the  plant  from  eight  to  twelve  hours. 
The  precipitating  tank  is  so  constructed  as  to  cause  intimate  and  thorough 
mixing  of  the  chemicals  with  the  water.  Thus  the  acids  are  neutralized, 
and  the  scale-forming  substances  are  precipitated  by  being  changed  to 
insoluble  substances  which  sink  to  the  bottom  of  the  precipitating  tank 
whence  they  are  readily  removed.  Some  of  the  lighter  substances 
remaining  in  suspension  are  carried  along  with  the  water  as  it  passes 
into  the  filters,  which  effectively  remove  all  suspended  matter.  This 
system  is  continuous  in  operation,  and  purification  is  accomplished 
without  appreciably  retarding  the  onward  flow  of  feed  water.  Fig.  237 
shows  a  modification  of  the  system.  The  chemicals  are  pumped  from 
the  "  chemical  tank  "  into  the  "  solution  tanks,"  where  the  feed  water 
and  chemical  solution  are  thoroughly  mixed.  The  treated  water  is  taken 
from  these  tanks  and  pumped  into  the  "  precipitating  tanks  "  where  a 
large  portion  of  the  scale-forming  element  is  precipitated.  From  the 
precipitating  tanks  the  water  is  forced  through  a  series  of  filters  to  the 
boiler. 

Fig.  238  illustrates  the  We-Fu-Go  system  of  water  purification.     In 


FEED-WATER  PURIFIERS  AND  HEATERS 


481 


PRECIPITATING 
TANK? 


FIG.  237.    General  Arrangement  of  Scaife  System  of  Feed-Water  Purification. 


I      1 


FIG.  238.   General  Arrangement  of  We-Fu-Go  System  of  Feed- Water  Purification. 


482 


STEAM   POWER  PLANT  ENGINEERING 


this  installation  the  water  supply  first  enters  the  settling  or  treating 
tanks  into  which  the  chemicals  are  fed.  A  thorough  mixture  is  effected 
by  the  use  of  the  two  armed  paddles  located  near  the  bottom  of  the 
tanks.  From  the  treating  tanks  the  water  flows  by  gravity  into  the 
filters,  which  remove  all  remaining  impure  solid  matter  which  does  not 
settle  to  the  bottom  of  the  treating  tank.  The  pipes  conducting  the 
water  from  the  settling  tanks  to  the  filter  are  fitted  with  a  flexible  joint 
and  float  so  that  the  outlets  are  near  the  surface  at  all  times,  rising  and 
falling  with  the  water  level.  From  the  filters  the  purified  water  gravi- 
tates into  the  clear  water  storage  reservoir,  from  which  it  is  pumped 
into  an  open  heater  and  thence  to  the  boiler.  This  system  is  intermit- 
tent in  operation,  and  in  order  to  provide  sufficient  time  for  thorough 
chemical  treatment  of  large  quantities,  two  or  more  settling  tanks  are 
employed.  Both  the  We-Fu-Go  and  Scaife  systems  are  modified  in  a 
number  of  ways  to  meet  different  conditions. 


FIG.  238a.    Anderson  System  for  Preventing  Corrosion  in  Condensers. 

Fig.  238a  shows  the  general  arrangement  of  the  Anderson  system 
for  preventing  corrosion  in  condensers  and  removing  oil  from  condensed 
steam.  The  method  consists  in  injecting  into  the  exhaust  steam  as  it 
passes  from  the  preheater  to  the  condenser  a  solution  containing  a 
coagulant  which  changes  the  emulsion  of  the  cylinder  oil  to  a  flaky 
condition  so  that  it  may  be  separated  by  settling,  flotation,  or  filtering. 
The  air  pump  delivers  the  water  to  the  settling  tank  F,  whence  it  is 
taken  to  the  open  gravity  filters  G,  G,  of  a  superficial  area  proportional 
to  the  amount  of  water  to  be  passed  and  containing  a  filter  bed  of  four 
feet  of  crushed  quartz.  This  will  run  about  four  days  without  any 
marked  difference  in  efficiency,  after  which  time  the  bed  is  stirred  to  a 
depth  of  two  feet  by  mechanical  agitators  and  flushed  with  clean  water, 
by  which  all  impurities  are  carried  to  the  sewer.  The  solution  is  pre- 


FEED-WATER  PURIFIERS  AND  HEATERS  483 

pared  in  tank  A,  in  which  the  water  level  is  preserved  by  a  ball  float 
and  into  which  filtered  water  is  admitted  through  pipe  B,  while  the 
substance  with  which  the  water  is  treated  is  pumped  in  through  the 
pipe  D  by  a  small  pump  operated  from  the  main  engine.  The  flow  to 
the  "  rose  head  "  above  the  condenser  is  controlled  by  the  valve  E, 
and  a  meter  in  this  pipe  records  the  amount  being  fed.  The  water 
ordinarily  required  for  "  make  up  "  is  sufficient  to  carry  in  the  solution. 
There  is  very  little  loss  of  water,  and  the  rapid  corrosion  of  the  con- 
denser tubes,  which  has  been  so  great  an  obstacle  to  the  successful  use 
of  surface  condensers,  is  much  reduced.  The  chemicals  used  perform 
a  twofold  duty,  viz.,  to  neutralize  the  water  and  make  it  chemically 
inactive  and  to  coagulate  the  oily  matter  contained  in  the  steam  so  that 
mechanical  nitration  is  possible.  (Power,  June  1903,  p.  304.) 

Water-softening  plants  cost  from  $4  to  $5  per  horse  power  for 
plants  of  1000  horse  power  and  less,  from  $3  to  $4  for  plants  of 
1000  to  2000  horse  power,  and  as  low  as  $1.50  for  plants  of  5000 
horse  power  or  more.  The  depreciation  of  wooden  tanks  is  as  high  as 
15  per  cent  a  year,  while  that  of  steel  tanks  should  not  be  greater  than 
5  per  cent.  Unless  wooden  tanks  are  considerably  cheaper  than  steel 
tanks  they  are  not  a  good  investment.  The  cost  of  water  purification 
varies  from  a  fraction  of  a  cent  to  2  cents  per  1000  gallons,  depending 
upon  the  size  of  the  plant  and  the  quantity  and  character  of  the  impuri- 
ties. (American  Electrician,  March,  1905,  p.  125.) 

Feed  Water  Purification:  Am.  Elecn.,  March,  1900,  p.  145,  April,  1900,  p.  190, 
Dec.,  1904,  p.  618;  Cassier's  Mag.,  April,  1904,  p.  506;  Engng.,  Oct.  25,  1901,  p.  595; 
Eng.  News,  May  22,  1902,  p.  408;  Eng.  Rec.,  April  5,  1902,  p.  322,  June  10,  1905; 
Jour.  Amer.  Chem.  Soc.,  Nov.,  1893,  p.  610;  Jour.  Soc.  Chem.  Ind.,  Aug.,  1901, 
p.  828;  Power,  Nov.,  1900,  p.  7,  Sept.,  1902,  p.  33,  Nov.,  1904,  p.  693;  R.R.  Gaz.,  Aug. 
24,  1900,  p.  568;  Elec.  Rev.,  Nov.  12,  1904;  Engr.  U.S.,  Oct.  15,  1903,  Jan.  1,  1906; 
Elec.  World,  Sept.  1,  1906. 

Water-Softening  Plants.  —Water-Softening  Processes:  Prac.  Erigr.  U.S., Mar  ,  1910. 
Four  Systems  of  Softening  Water  for  Industrial  Purposes:  Eng.  News,  July  2,  1903, 
p.  4.  An  Inquiry  into  the  Working  of  Various  Water  Softeners:  Inst.  of  Mech.  Engrs., 
Dec.,  1903.  Report  on  Soft  Water  for  Locomotive  Plants:  Eng.  News,  March  17, 
1904.  The  Development  of  Water  Purification  in  the  U.S.:  R.R.  Gaz.,  38-19.  Gen- 
eral Information  on  Water  Softening:  Eng.  News,  May  26,  1904,  p.  500,  508,  June 
2,  1904,  p.  530;  Eng.  Rec.,  Oct.  24,  1903,  p.  483;  Loco.  Engng.,  Nov.,  1903,  p.  501; 
Elec.  Engr.,  Lond.,  April  21,  1905;  Ir.  and  Coal  Tr.  Rev.,  Sept.  1,  1905;  Jour.  W.  Soc. 
Engrs.,  Dec.,  1905;  Eng.  and  Min.  Jour.,  Dec.  2,  1905;  Am.  Engr.  and  R.R.  Jour., 
Jan.,  1905. 

Harris-Anderson  Water  Softener:   Engng.,  Aug.  15,  1902,  July  10,  1903. 

Kennicott  Water  Softeners:  Am.  Elecn.,  Nov.,  1902,  p.  545;  Eng.  News,  May  15, 
1902,  p.  386;  Eng.  Rec.,  May  3,  1902,  p.  419;  St.  Ry.  Jour.,  April  2,  1904, 
p.  545. 


484  STEAM   POWER   PLANT  ENGINEERING 

Burt  Continuous  Water- Softening  Process:  Eng.  News,  Sept.,  15,  1904,  p.  238; 
Engr.  U.S.,  July  15,  1905,  p.  426. 

Bachman  Method  of  Water  Purification:   St.  Ry.  Review,  May  15,  1900,  p.  282. 

We-Fu-Go  and  Scaife  Systems:  St.  Ry.  Rev.,  Oct.,  1901,  p.  771;  Engr.  U.S., 
Jan.  1,  1903,  p.  90. 

Holmes  System  of  Water  Purification:   Power,  April,  1905,  p.  248. 

The  American  Water  Purifier  and  Softener:   Eng.  U.S.,  Aug.  1,  1904,  p.  551. 

248.  Economy  of  Preheating  Feed  Water.  —  Although  a  feed  water 
heater  acts  to  some  extent  as  a  purifier  its  primary  function  is  that  of 
heating  the  feed  water.  Generally  speaking,  for  every  10  degrees 
that  the  feed  water  is  heated  there  is  a  gain  in  heat  of  1  per  cent  and  a 
corresponding  saving  of  coal,  if  the  heat  which  warms  the  feed  water 
would  otherwise  be  wasted.  Again,  the  smaller  the  difference  in 
temperature  between  the  steam  and  the  feed  water  the  less  will  be  the 
strain  on  the  boiler  shell  due  to  unequal  expansion  and  contraction,  an 
item  of  no  small  consequence. 

If  A  represents  the  total  heat  of  one  pound  of  steam  above  32 
degrees  F.,  £0  the  temperature  of  the  cold  water,  and  t  the  temperature 
of  the  water  leaving  the  heater,  then  S,  the  per  cent  gain  in  heat  due 
to  heating  the  feed  water,  may  be  expressed 

s  =  100     ('7°-^  '  (104) 

A  —  (fQ  —  6Z) 

The  expression  is  not  theoretically  correct,  since  it  assumes  a  con- 
stant value  of  unity  for  the  specific  heat,  whereas  the  specific  heat 
varies  with  the  temperature.  The  variation  is  so  slight,  however, 
that  it  may  be  neglected  for  all  practical  purposes. 

Example:  Steam  pressure  100  pounds  gauge;  temperature  of  water 
entering  heater  80  degrees  F.;  temperature  of  water  leaving  heater 
210  degrees  F.  Required,  saving  due  to  heating  the  feed  water. 

Here  X  (from  steam  tables)  is  1185,  t0  =  80,  t  =  210. 

(210-80) 


11 85 -(80 -32) 
=  11.42  per  cent. 

This  formula  gives  the  thermal  saving  only,  and  the  first  cost  of  the 
heater,  interest,  depreciation,  attendance,  and  repairs  must  be  taken 
into  consideration  before  the  net  saving  measured  in  dollars  and  cents 
is  ascertained.  In  the  average  installation  the  net  saving  is  a  sub- 
stantial one. 


FEED-WATER  PURIFIERS  AND  HEATERS 


485 


62  based  upon  formula  (104)  may  be  used  m  determining  the 
percentages  of  saving  due  to  the  increase  in  feed-water  temperature. 

TABLE   62. 

PERCENTAGE  OF  SAVING  FOR  EACH  DEGREE  OF  INCREASE  IN  TEMPERATURE 

OF    FEED   WATER. 


Initial 
Temp, 
of  Feed. 

Boiler  Pressure  above  Atmosphere. 

i 

0 

20 

40 

60 

80 

100 

120 

140 

160 

180 

200 

32 

.0872 

.0861 

.0855 

.0851 

.0847 

.0844 

.0841 

.0839 

.0837 

.0835 

.0833 

40 

.0878 

.0867 

.0861 

.0856 

.0853 

.0850 

.0847 

.0845 

.0843 

.0841 

.0839 

50 

.0886 

.0875 

.0868 

.0864 

.0860 

.0857 

.0854 

.0852 

.0850 

.0848 

.0846 

60 

.0894 

.0883 

.0876 

.0872 

.0867 

.0864 

.0862 

.0859 

.0856 

.0855 

.0853 

70 

.0902 

.0890 

.0884 

.0879 

.0875 

.0872 

.0869 

.0867 

.0864 

.0862 

.0860 

—80 

.0910 

.0898 

.0891 

.0887 

.0883 

.0879 

.0877 

.0874 

.0872 

.0870 

.0868 

90 

.0919 

.0907 

.0900 

.0895 

.0888 

.0887 

.0884 

.0883 

.0879 

.0877 

.0875 

100 

.0927 

.0915 

.0908 

.0903 

.0899 

.0895 

.0892 

.0890 

.0887 

.0885 

.0883 

110 

.0936 

.0923 

.0916 

.0911 

.0907 

.0903 

.0900 

.0898 

.0895 

.0893 

.0891 

120 

.0945 

.0932 

.0925 

.0919 

.0915 

.0911 

.0908 

.0906 

.0903 

.0901 

.0899 

130 

.0954 

.0941 

.0934 

.0928 

.0924 

.0920 

.0917 

.0914 

.0912 

.0909 

.0907 

140 

.0963 

.0950 

.0943 

.0937 

.0932 

.0929 

.0925 

.0923 

.0920 

.0918 

.0916 

150 

.0973 

.0959 

.0951 

.0946 

.0941 

.0937 

.0934 

.0931 

.0929 

.0926 

.0924 

160 

.0982 

.0968 

.0961 

.0955 

.0950 

.0946 

.0943 

.0940 

.0937 

.0935 

.0933 

170 

.0992 

.0978 

.0970 

.0964 

.0959 

.0955 

.0952 

.0949 

.0946 

.0944 

.0941 

180 

.1002 

.0988 

.0981 

.0973 

.0969 

.0965 

.0961 

.0958 

.0955 

.0953 

.0951 

190 

.1012 

.0998 

.0989 

.0983 

.0978 

.0974 

.0971 

.0968 

.0964 

.0962 

.0960 

200 

.1022 

.1008 

.0999 

.0993 

.0988 

.0984 

.0980 

.0977 

.0974 

.0972 

.0969 

210 

.1033 

.1018 

.1009 

.1003 

.0998 

.0994 

.0990 

.0987 

.0984 

.0981 

.0979 

220 

.1029 

.1019 

.1013 

.1008 

.1004 

.1000 

.0997 

.0994 

.0991 

.0989 

230 

.1039 

.1031 

.1024 

.1018 

.1012 

.1010 

.1007 

.1003 

.1001 

.0999 

240 

.1050 

.1041 

.1034 

.1029 

.1024 

.1020 

.1017 

.1014 

.1011 

.1009 

250 



.1062 

.1052 

.1045 

.1040 

.1035 

.1031 

.1027 

.1025 

.1022 

.1019 

Multiply  the  factor  in  the  table  corresponding  to  any  given  initial  temperature  of  feed  water 
and  boiler  pressure  by  the  total  rise  in  feed-water  temperature;  the  product  will  be  the  percent- 
age of  saving. 

Feed  Water  Heating.  —  How  Should  Feed  Water  be  Heated  ?  —  Power,  July,  1907, 
p.  456;  Feed  Water  Heating:  Engr.  U.S.,  Jan.  1,  1906,  p.  8,  Aug.  15,  1904,  p.  15; 
St.  Ry.  Jour.,  July  22,  1905,  p.  145;  Am.  Elecn.,  Dec.,  1904,  p.  570;  Am.  Elecn., 
Nov.,  1904;  Engr.,  Lond.,  July  28,  1905. 

249.  Classification  of  Feed- Water  Heaters.  —  Feed-water  heaters 
may  be  classified  according  to  the  source  of  heat,  as 

1.  Exhaust  steam,  in  which  the  heat  is  received  from  the  exhaust  of 
engines,  pumps,  etc. 

2.  Flue  gas,  in  which  the  waste  chimney  gases  are  the  source  of  the 
heat. 

3.  Live  steam  purifiers,  or  those  using  steam  at  boiler  pressures;  or 
according  to  the  method  of  heat  transmission,  as 


486 


STEAM  POWER  PLANT  ENGINEERING 


1.  Open  heaters,  in  which  the  steam  and  feed  water  mingle  and  the 
steam  in  condensing  gives  up  its  heat  directly  to  the  water. 

2.  Closed  heaters,  in  which  the  steam   and  water  are  in  separate 
chambers  and  the  steam  gives  up  its  heat  to  the  water  by  conduction. 

Heaters  may  also  be  classified  according  to  the  pressure  of  the  heat- 
ing steam,  as 

1.  Vacuum  or  primary,  in  which  the  pressure  is  less  than  atmos- 
pheric and  applies  particularly  to  heaters  utilizing  the  exhaust  of  con- 
densing engines.     These  are  always  of  the  closed  type.     Open  heaters 
in  which  the  pressure  is  less  than  atmospheric  are  not  usually  classed 
as  vacuum  heaters. 

2.  Atmospheric  or  secondary,  in  which  the  pressure  is  atmospheric 
or,  literally,  that  corresponding  to  the  back  pressure  on  the  engines 
and  pumps. 

3.  Pressure,  in  which  the  pressure  corresponds  to  that  in  the  boiler 
and  in  which  the  heat  is  used  primarily  for  purifying  purposes. 

Heaters  may  be  still  further  classified  as 

1.  Induced,  in  which  only  such  steam  is  admitted  as  is  induced  by 
its  condensation.     That  is,  the  feed  water  condenses  the  steam.     This 
creates  a  partial  vacuum  which  draws  in  more  steam. 

2.  Through,  in  which  all  the  steam  is  forced  through   the  heater 
irrespective  of  condensation. 

CLASSIFICATION  OF  A  FEW  TYPICAL  HEATERS. 


Exhaust  steam 


Open Atmospheric. 


,      ( Atmospheric 

—  -' '   ( Vacuum  or  pressure 


Flue  Gas 

Live  Steam Open Pressure. 


Cochrane 

Hoppes 

Stillwell 

Webster 

Wainwright)  Water 

Wheeler...)  Tube 

Otis j  Steam 

Berryman  .JTube 
Green 
American 
Sturtevant 
Hoppes 
Baragwanath 


250.  Open  Heaters.  —  Fig.  239  gives  a  sectional  view  of  a  Cochrane 
special  feed  heater  and  receiver  and  is  a  typical  example  of  an  open 
heater.  Exhaust  steam  enters  the  heater  through  a  fluted  oil  separa- 
tor as  indicated,  and  passes  out  at  the  top,  while  the  oily  drips  are 
automatically  drained  to  waste  by  a  suitable  ventilated  float.  The 
feed  water  enters  through  an  automatic  valve  and  is  distributed  over 


PEED-WATER  PURIFIERS  AND  HEATERS 


487 


a  series  of  copper  trays  so  arranged  and  constructed  that  the  water  is 
forced  to  fall  in  a  finely  divided  stream  before  reaching  the  reservoir  in 
the  bottom.  The  steam  coming  in  contact  with  the  water  particles 
gives  up  latent  heat  and  condenses.  Much  of  the  scale-forming  ele- 
ment is  deposited  on  the  surface  of  the  trays,  from  which  it  is  readily 
removed.  The  suspended  matter  is  eliminated  by  a  coke  filter  in  the 


FIG.  239.    Cochrane  Special  Heater  and  Receiver. 

bottom  of  the  chamber,  and  the  floating  impurities  are  decanted  by  a 
skimmer  or  overflow  weir.  The  particular  heater  shown  in  the  illustra- 
tion is  especially  designed  for  use  in  a  steam-heating  plant;  i.e., 
besides  performing  all  the  functions  of  an  open  heater,  it  provides  for 
the  reception  and  heating  of  the  condensation  returned  to  it  from  the 
heating  system. 


488 


STEAM  POWER  PLANT  ENGINEERING 


Fig.  240  gives  a  sectional  view  of  a  Webster  "  star  vacuum  "  heater. 
Water  enters  the  heater  through  balanced  valve  F,  which  is  controlled 
by  float  E,  and  is  deflected  over  a  series  of  perforated  copper  trays  T,  T. 
Exhaust  steam  enters  at  A,  passes  through  oil  filter  S,  and,  mingling 
with  the  finely  divided  streams  of  water,  gives  up  its  latent  heat  and  is 
condensed.  Only  so  much  steam  enters  the  heater  as  is  condensed  by 
the  feed  water.  The  condensed  steam  and  feed  water  fall  to  the  bottom 


FIG.  240.    Section  Through  Webster  Heater. 

of  the  upper  chamber,  maintaining  a  practically  constant  level  WW . 
From  this  upper  or  heater  chamber  the  water  gravitates  to  the  settling 
chamber  at  the  bottom,  through  down-cast  pipe  CB.  From  the  set- 
tling chamber  the  water  rises  through  perforated  screen  M  and  filtering 
material  P  to  the  outlet  0.  A  large  portion  of  the  scale-forming  ele- 
ment is  precipitated  on  the  trays  or  collects  in  the  settling  chamber  at 
the  bottom. 


FEED-WATER  PURIFIERS  AND  HEATERS 


489 


Fig.  241  shows  a  section  through  a  Hoppes  open  heater,  illustrating 
the  "  pan  "  type.  Exhaust  steam  enters  at  H,  passes  through  oil  filter  0, 
and  completely  surrounds  pans  T,  T.  The  feed  water  enters  at  B, 
and  the  rate  of  flow  is  regulated  by  valve  F,  which  is  controlled  by  a 


FIG.  241.    Hoppes  Horizontal  Feed- Water  Heater. 

suitable  float  in  the  lower  part  of  the  chamber.  The  water  in  flowing 
over  the  sides  and  bottoms  of  the  pans  comes  in  direct  contact  with 
the  steam, 

251.  Combined  Open  Heater  and  Chemical  Purifier.  —  Combined  feed- 
water  heaters  and  chemical  purifiers  are  finding  increased  favor  with 
engineers  in  many  districts  where  the  feed  water  is  particularly  bad. 
A  description  of  the  Webster  combination  will  be  found  in  Part  II  of 
the  general  catalogue  issued  by  the  Warren  Webster  Company,  Camden, 
N.J.     A  description  of  the  Cochrane-Sorge  combined  heater  and  chem- 
ical purifier  will  be  found  in  the  heater  catalogue  issued  by  the  Harrison 
Safety  Boiler  Works,  Philadelphia,  Pa. 

252.  Temperatures  in  Open  Heaters.  —  The  temperature  to  which 
feed  water  is  raised  in  an  open  heater  may  be  determined  as  follows : 

Let  A  represent  the  total  heat  of  steam  corresponding  to  the  pressure 

in  the  heater, 

tQ  the  temperature  of  the  water  entering  heater, 
t  the  temperature  of  the  water  leaving  heater,  and 
S  the  ratio  of  exhaust  steam  to  the  feed  water,  by  weight. 


490 


STEAM  POWER  PLANT  ENGINEERING 


Then,  allowing  a  loss  of  10  per  cent  due  to  radiation,  etc., 
0.9  S  (A  - 1  +  32)  will  be  the  B.T.U.  given  up  by  the  exhaust  steam  to 
each  pound  of  feed  water,  and  (t  —  tQ)  will  be  the  B.T.U.  absorbed  by 
each  pound  of  water. 

Therefore  0.9  S  (I  -  t  +  32)  =  t  -  t0,  from  which 

^MsM+m.  (105) 


t  = 


1  +  0.9 


If  more  steam  passes  through  the  heater  than  can  be  condensed  by 
the  feed  water,  then  this  equation  gives  t  a  fictitious  value;  in  other 
words,  t  can  never  be  greater  than  the  temperature  of  the  exhaust 
steam. 

Substituting  t  =  212,  the  maximum  obtainable  temperature  with 
exhaust  steam  at  atmospheric  pressure,  and  solving  for  S,  we  find  that 
only  17  per  cent  of  the  main  engine  exhaust  is  necessary  to  heat  the 
feed  water  to  a  maximum.  t0  is  assumed  to  be  60  degrees  F. 

Table  63  has  been  determined  from  this  equation  and  gives  the  final 
temperatures  obtainable  in  open  heaters  for  various  conditions  of 
operation. 

TABLE   63. 
FINAL    FEED-WATER    TEMPERATURES.     OPEN    HEATER. 

(Temperature  of  steam,  212  degrees  F.) 


Initial  Temperature  of  Feed  Water,  Degrees  F. 

40 

50 

60 

70 

80 

90 

100 

110 

120 

130 

1 

2 

60.1 

69.9 

79.7 

89.5 

94.4 

109.2 

119.0 

128.8 

138.7 

148.5 

p 

3 

69.9 

79.6 

89.3 

90.1 

108.8 

118.6 

128.3 

138.0 

147.8 

157.5 

8 

4 

79.5 

89.1 

98.8 

108.5 

118.1 

127.8 

137.4 

147.1 

156.7 

166.4 

§,  t 

5 

89.0 

98.5 

108.1 

117.7 

127.2 

136.8 

146.4 

155.9 

165.5 

175.1 

21 

6 

98.3 

107.7 

117.2 

126.7 

136.2 

145.7 

155.2 

164.7 

174.2 

183.6 

^    :3 

7 

107.4 

116.8 

126.2 

135.6 

145.0 

154.4 

163.8 

173.2 

182.5 

192.1 

o  ^ 
H  =| 

8 

116.4 

125.7 

135.0 

144.4 

153.7 

163.0 

172.4 

181.8 

191.0 

200.3 

*£ 

9 

125.2 

134.5 

143.7 

153.0 

162.2 

171.5 

180.7 

190.0 

199.2 

208.5 

«* 

10 

133.3 

143.1 

152.3 

161.4 

170.6 

179.8 

189.0 

198.1 

207.3 

212.0 

1 

11 

142.5 

151.6 

160.7 

169.7 

178.9 

188.2 

197.0 

206.2 

212.0* 

212.0* 

I 

12 

150.9 

159.9 

168.9 

177.9 

187.0 

196.0 

205.0 

212.0* 

212.0* 

212.0* 

*  All  of  the  steam  not  condensed. 

Example  :  A  power  plant  has  1200  I.H.P.  of  engines  using  20  pounds 
of  steam  per  I."H.P.  hour.  Auxiliaries  use  equivalent  of  10  per  cent  of 
main  engine  steam.  Pressure  in  heater  0  pounds  gauge,  temperature 
of  hot- well  supply  110  degrees  F.  Required  temperature  of  feed  water 
leaving  heater. 


FEED-WATER  PURIFIERS  AND  HEATERS  491 

Here         X  =  1146  (from  steam  tables),  tQ  =  110,  S  =  0.10. 

Substituting  these  values  in  (105), 

0.9  X  0.10  (1146  - 1  +  32)  =  t  -  110. 

t  =  198  degrees  F. 

253.  Pan  Surface  Required  in  Open  Feed- Water  Heaters.  —  Pan  or 

tray  surface  required  varies  according  to  the  quality  of  the  water  with 
regard  to  both  scale-making  material  and  mud,  and  may  be  approxi- 
mated by  the  formula 

r»  r  ,,  Lb.of  water  heated  per  hr.X  horsepower  /,n(-  N 
Pan  surface,  sq.  ft.  = c K .  (105a) 


Vertical  Type. 

Horizontal 
Type. 

For  very  muddy  water  c                    

118 

110 

Slightly  muddy  water  c            

166 

155 

For  clean  water  c                    

500 

400 

254:.  Size  of  Shell,  Open  Heaters.  —  General  proportions  of  open 
heaters  vary  considerably  on  account  of  the  different  arrangements  of 
pans  or  trays,  filter  and  oil-extracting  devices.  A  fair  idea  of  the  size 
of  shell  required  may  be  obtained  by  the  formulas 

.   ,    n        Horse  power  /i/us\ 

Area  of  shell  = : fFi — 7—7  •  (106) 

a  X  length  in  feet 

T       ii     *   i    n  Horse  power  /i/vr\ 

Length  of  shell  = : — K — r  *  (107) 

a  X  area  in  square  feet. 

a  =  2.15  for  very  muddy  water. 
a  =  6       for  slightly  muddy  water. 
a  =  8       for  clean  water. 

The  horse  power  in  this  case  is  obtained  by  dividing  the  weight  of  water 
heated  per  hour  by  the  steam  consumption  of  the  engine  per  horse 
power  per  hour. 

Pans  containing  2.5  square  feet  and  less  are  usually  made  round,  and 
larger  sizes  rectangular  in  plan.  When  circumstances  will  permit  it 
is  better  to  have  not  more  than  six  pans  in  any  one  tier,  since  it  is 
advisable  to  proportion  the  pans  so  as  to  obtain  as  low  a  velocity  over 
each  as  practicable. 

Distance  between  trays  or  pans  is  seldom  less  than  one-tenth  the 
width  for  rectangular  and  one-fourth  the  diameter  for  round  pans. 
Volume  of  storage  and  settling  chamber  in  horizontal  heaters  varies 


492 


STEAM  POWER  PLANT  ENGINEERING 


from  0.25  for  good  quality  of  water  to  0.4  of  the  volume  of  the  shell 
for  muddy  water,  0.33  being  about  the  average.  In  the  vertical  type 
the  settling  chamber  represents  respectively  0.4  and  0.6  the  volume  of 
the  shell  with  clear  and  muddy  water.  Filters  occupy  from  10  to  15 
per  cent  of  the  volume  of  the  shell  in  the  horizontal  type  and  from  15 
to  20  per  cent  in  the  vertical  type,  the  smaller  percentage  corresponding 
to  clear  water  and  the  larger  to  muddy  water  or  water  containing  a  con- 
siderable quantity  of  impurities. 

Open  Heaters:   Cassier's  Mag.,  Aug.,  1903,  p.  33;  Engr.  U.S.,  Jan.  1,  1906,  pp.  17, 
78;  St.  Ry.  Jour.,  Feb.  4,  1905,  p.  227;  Am.  Elecn.,  Sept.,  1905,  p.  481. 


SURFACE    BLOW 


IiXHAUST  FROM 

HEATER 


FIG.  242.     Goubert   Single-Flow    Closed 
Heater. 


255.  Classification  of  Closed 
Heaters.  —  Closed  heaters  may 
be  grouped  into  two  classes: 

1.  Water  tube,  Fig.  242,  and 

2.  Steam  tube,  Fig.  246. 
Closed    heaters,     both    water 

tube  and  steam  tube  may  oper- 
ate with: 

1.  Parallel  currents,  where  the 
water    and    steam   flow   in    the 
same  direction,  Fig.  242,  or  with 

2.  Counter  currents,  where  the 
water  and  steam  flow  in  opposite 
directions,  Fig.  244. 


FIG.  243.     Details  of  Expansion  Joint, 
Goubert  Heater. 


Water-tube  heaters  may  be  still  further  classified  as 
1.  Single-flow,  in  which  the  water  flows  through  the  heaters  in  one 
direction  only,  Fig.  242. 


FEED-WATER  PURIFIERS  AND  HEATERS 


493 


2.  Multi-flow,  in  which  the  water  flows  back  and  forth  a  number  of 
times,  as  in  Fig.  244. 

3.  Coil  heater,  in  which  the  water  flows  through  one  or  more  coils, 
as  in  Fig.  245. 

256.  Water-Tube,    Closed    Heaters.  —  Fig.    242    shows    a    section 
through  a  feed-water  heater  of  the  single-flow  straight-tube  type.     The 


FIG.  244.     Wainwright  Multi-Flow 
Closed  Heater. 


FIG.  245.     Typical  Coil  Heater. 


tubes  are  of  plain  brass  and  the  shell  of  cast  iron.  The  tubes  are 
expanded  into  the  tube  sheets  by  a  roller  expander.  To  provide  for 
expansion  the  upper  tube  sheet  and  water  chamber  are  secured  to  the 
main  shell  by  means  of  a  special  expansion  joint  the  details  of  which 
are  shown  in  Fig.  243.  R  is  a  ring  or  gasket  of  soft  annealed  copper 


494  STEAM  POWER  PLANT  ENGINEERING 

and  Cr,  G  two  gaskets  of  special  packing  with  brass  wire  cloth  insertion. 
These  gaskets  form  a  flexible  expansion  joint  between  C  and  tube 
sheet  D,  so  that  the  whole  upper  chamber,  which  is  carried  solely  by 
the  tubes,  is  free  to  move  up  and  down  as  the  tubes  expand  or  contract 
under  varying  temperatures. 

Fig.  244  shows  a  section  through  a  Wainwright  heater,  illustrating 
the  multi-flow  water-tube  type.  The  body  of  the  heater  is  of  cast 
iron,  the  tubes  of  corrugated  copper.  The  water  passes  through 
the  tubes  and  the  steam  surrounds  them.  The  feed  water  and 
exhaust  steam  do  not  mingle,  and  hence  the  oil  in  the  exhaust 
does  not  contaminate  the  water.  The  water  chambers  are  divided 
into  several  compartments,  as  shown  in  the  illustration,  and  the  par- 
titions are  so  arranged  that  the  flow  of  feed  water  is  directed  back 
and  forth  through  the  various  groups  of  tubes  in  succession.  This 
arrangement  gives  a  higher  velocity  of  flow  than  the  non-return 
type  of  heater,  and  therefore  increases  the  rate  of  heat  absorp- 
tion. The  mud  and  impurities  settle  at  the  bottom  and  are 
discharged  through  the  mud  blow-off.  Such  impurities  as  rise  to 
the  surface  are  removed  by  the  surface  blow-off.  The  tubes  are  cor- 
rugated to  allow  for  expansion  and  at  the  same  time  to  increase  the 
transmission  of  heat.  Referring  to  Fig.  244:  Exhaust  steam  enters 
at  A  and  leaves  at  E,  and  the  portion  which  is  condensed  is  drawn  off 
at  D.  Feed  water  enters  at  I  and  is  discharged  at  0.  P,  P  are  mud 
blow-offs  and  S  is  an  opening  for  a  safety  valve.  Table  66  gives  results 
of  tests  showing  the  relative  efficiencies  of  plain  and  corrugated  tubes 
for  various  velocities. 

Fig.  245  shows  a  partial  section  through  a  Harrisburg  feed-water 
heater.  This  apparatus  is  a  typical  example  of  the  coiled-tube  heater. 
Three  sets  of  concentric  copper  coils  are  brazed  to  gun-metal  manifolds 
and  supported  by  clamp  stays  as  indicated  in  the  illustration.  Feed 
water  enters  the  heater  at  the  bottom  manifold  and  passes  through 
the  coils  to  the  feed  outlet.  The  exhaust  steam  enters  the  heater  at 
the  bottom  and  surrounds  the  coils  in  its  passage  to  the  outlet  at  the 
top.  The  coils  are  designed  to  withstand  a  pressure  of  600  pounds 
per  square  inch. 

257.  Steam-Tube,  Closed  Heaters.  —  Fig.  246  shows  a  section 
through  an  Otis  heater,  illustrating  the  steam-tube  type.  Here  the 
exhaust  steam  passes  through  the  tubes  which  are  surrounded  by  the  feed 
water.  The  exhaust  steam  enters  at  A,  and  passes  down  one  section 
of  tubes  into  the  enlarged  space  of  the  water  and  oil  separator  0,  in 
which  the  condensation  and  oil  are  deposited.  From  this  chamber  the 
steam  passes  up  through  the  other  section  of  tubes  to  outlet  C,  thus 


FEED-WATER  PURIFIERS  AND  HEATERS 


495 


passing  twice  through  the  entire  length  of  the  heater.  The  water  enters 
at  E  and  is  discharged  at  G.  R  is  the  blow-off  opening.  The  tubes 
are  of  seamless  brass  and  are  curved  to  allow  for  expansion.  Condensed 
steam  is  withdrawn  at  P. 

Fig.  247   shows  a  partial   section  through  a  Baragwanath   steam 
jacketed  steam-tube  heater.     Exhaust  steam  enters  at  A,  passes  up 


p 


FIG.  246.     Otis  Steam-Tube  Feed-Water 
Heater. 


FIG.  247.    Baragwanath  Steam-Jacketed 
Feed-Water  Heater. 


through  the  tubes,  returns  down  annular  space  E  between  the  inner 
shell  and  jacket,  and  passes  out  at  B.  Feed  water  enters  at  C  and 
leaves  at  D.  E  is  the  scum  blow-off,  G  the  heater  drain,  and  H  the 
jacket  drain. 


496  STEAM  POWER  PLANT  ENGINEERING 

258.  Heating  Surface,  Closed  Heaters.  —  It  is  generally  assumed 
that  the  transfer  of  heat  between  two  bodies  is  directly  proportional 
to  the  difference  in  temperature  between  them. 

Let    TQ  =  temperature  of  the  water  entering  the  heater. 
T2  =  temperature  of  the  water  leaving  the  heater. 
Tg  =  temperature  of  the  exhaust  steam. 
A  =  square  feet  of  transmitting  surface. 
T  =  temperature  of  a  unit  of  water  t  seconds  after  entering 

the  heater. 
h   =  B.T.U.  absorbed  per  square  foot  per  second  per  degree 

difference  in  temperature  between  the  steam  tempera- 

ture Ts  and  the  water  temperature  T. 
t  =  time  in  seconds. 
w  =  number  of  pounds  of  feed  water  per  second. 

A 

Then  —  =  square  feet  of  surface  brought  in  contact  with  one  pound  of 

water  per  second, 

and  dT,  the  rate  at  which  the  temperature  of  the  water  is  increasing 
at  this  instant,  will  be 

dT=  —  (T,-T)dt.  (108) 


--""  (109) 


Integrating, 


<"» 


Let  W  =  number  of  pounds  of  feed  water  heated  per  hour. 

U  =  B.T.U.  transmitted  to  the  feed  water  per  square  foot  of  sur 

face  per  hour  per  degree  difference  in  temperature. 
Then  (112)  may  be  written 


from  which 


FEED-WATER  PURIFIERS  AND  HEATERS  497 

Knowing  the  weight  of  water  to  be  heated,  the  temperature  of  the 
s^eam,  the  desired  temperature  of  the  feed  water,  and  the  coefficient  of 
heat  transmission,  U,  this  equation  enables  one  to  determine  the  area 
of  heating  surface  required  for  the  given  conditions.  Since  the  extent 
of  heating  surface  increases  rapidly  as  T2  approaches  Ts,  and  becomes 
infinity  for  T2  =  T8,  it  is  desirable  to  limit  T2  to  some  practical  figure. 
An  average  maximum  for  T2  —  T8  —  4. 

Table  64  has  been  calculated  from  this  formula  and  gives  the  square 
feet  of  heating  surface  necessary  to  heat  1000  pounds  of  water  per  hour 
for  different  ranges  in  temperature. 

Mean  Temperature  Difference. 

If  we  let  d  =  average  temperature  difference  between  the  steam 
and  feed  water,  then 

A  Ud  =  heat  given  out  by  the  steam  per  hour. 
W  (T2  —  T0)  =  heat  absorbed  by  the  feed  water  per  hour. 
AUd  =  W(T2-T0).  (115) 

d  =  W  (Tl~T^  -  (116) 

A.  (J 

ATT  T        T 

From  (113),  2$L  =  log,  g^ ,  (117) 

Therefore  d  =      T*~T       -  (118) 

Table  65  has  been  calculated  from  formula  (118)  and  gives  the 
mean  temperature  difference  for  various  conditions  of  operation. 

The  arithmetic  mean  temperature  difference  dt  may  be  taken  with 
safety  for  the  average  heater  problem  and  has  the  advantage  of  sim- 
plicity. 

d,=  T8-  T«+T?»  (119) 

£ 

Closed  heaters  are  sometimes  rated  on  the  basis  of  \  square  foot  of 
heating  surface  per  horse  power,  i.e.,  a  heater  with  500  square  feet  of 
heating  surface  would  be  rated  at  1500  horse  power. 

259.  Heat  Transmission  in  Closed  Heaters.  —  Table  66  gives  the 
results  of  a  series  of  tests  on  the  absorption  of  heat  by  water  passing 
through  brass  and  copper  tubes  surrounded  by  steam.  The  curves  in 
Fig.  248  were  plotted  from  the  data  given  in  this  table.  An  inspection 
of  the  table  and  the  curves  will  show  that  the  absorption  of  heat  per 
square  foot  of  surface  per  degree  difference  in  temperature  varies  with 


498 


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STEAM   POWER  PLANT  ENGINEERING 


the  velocity  of  the  water  and  the  material  and  character  of  the  tubes. 
Increasing  the  velocity  of  the  water  passing  through  the  heater  in- 
creases the  rate  of  heat  transmission  and  thereby  renders  the  heating 

surface  more  effective.  In  order  to 
employ  moderately  high  velocities 
and  at  the  same  time  allow  suffi- 
cient time  in  which  to  raise  the 
temperature  to  a  maximum,  the 
tubes  should  be  as  long  as  prac- 
ticable and  of  small  diameters. 
Other  things  being  equal,  a  heater 
containing  a  large  number  of  tubes 
of  small  diameter  is  more  economical 
than  one  containing  a  small  number 
of  large  tubes.  It  is  important  to 
proportion  the  heater  according  to 
the  amount  of  water  to  be  heated 
and  the  maximum  temperature  to 
which  the  water  must  be  raised. 
In  designing  a  heater,  then,  the 
maximum  amount  of  heat  to  be 
transmitted  per  degree  difference  in 
temperature  per  hour  per  square  foot 
should  be  assumed,  and  the  velocity 
of  the  water  made  such  that  it  is 
capable  of  absorbing  this  amount. 
A  good  average  figure  for  multi- 
flow  heaters  is  U=  250  B.T.U.  for 
plain  brass  or  copper  tubes  and 
U=  300  B.T.U.  for  corrugated  tubes  with  a  water  velocity  of  50  feet 
per  minute;  for  single-flow  heaters,  17=  175  (for  plain  brass)  with  a 
water  velocity  of  12.5  feet  per  minute  and  for  coil  heaters  U=  300 
(copper)  with  a  water  velocity  of  150  feet  per  minute.  These  figures 
are  for  water-tube  heaters  only.  For  steam-tube  heaters  (iron  tubes) 
a  good  average  figure  is  U=  120. 

Experiments  show  that  heaters  and  condensers  operating  with 
counter-currents  are  more  efficient  and  are  capable  of  obtaining  a 
higher  final  temperature  than  those  operating  with  parallel  currents. 

Example:  Determine  the  size  of  vacuum  and  atmospheric  heaters 
for  a  condensing  plant  of  1200  I.H.P.  Engines  use  20  pounds  of 
steam  per  I.H.P.  hour;  auxiliaries  use  the  equivalent  of  10  per  cent 
of  the  main  engine  steam;  vacuum  25  inches  referred  to  30-inch 


50       75      100 
FIG.  248. 


FEED-WATER  PURIFIERS  AND  HEATERS  501 

barometer;    feed  water,   TQ  =  50  degrees;   temperature    of  hot   well, 
T2  =  110  degrees;  coefficient  of  heat  transmission,  U=  300  B.T.U. 

Vacuum  or  Primary  Heater. 
Feed  water  for  main  engines, 

20  X  1200  =  24,000  pounds  per  hour. 
Feed  water  used  by  auxiliaries, 

10  per  cent  of  24,000  =  2400  pounds  per  hour. 
Total  feed, 

W=  24,000  +  2400  =  26,400  pounds  per  hour. 
From  formula  (114), 


26,400  j 


300          "134-110 
=  110  square  feet. 

On  the  basis  of  J  square  foot  of  surface  per  horse  power  the  rating  of 
this  heater  will  be 

110  X  3  =  330  horse  power. 

Atmospheric  or  Secondary  Heater. 

The  temperature  of  the  feed  water  leaving  the  atmospheric  heater, 
formula  (105),  will  be 

,       t0  +  0.9  S(\  +  32) 

1  +  0.9  S 
where  S  =  .10,  t0  =  110  degrees,  X  =  1146  B.T.U. 

whence  t  =  11Q  +  Q-9  X  0.10  (1146  +  32) 

1  +  0.9  X  0.10 
=  198  degrees. 

The  required  surface  is 


where  Tt  =  212,  T0  =  110,  T2  =  198, 


whence         A  -  2^^  loge  212~110 
300         3C  212  -  198 

=  175  square  feet. 
The  horse-power  rating  will  be 
175  X  3  =  525  horse  power. 


502 


STEAM  POWER  PLANT  ENGINEERING 


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FEED-WATER  PURIFIERS  AND  HEATERS 


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504  STEAM  POWER  PLANT  ENGINEERING 

260.  Open  vs.  Closed  Heaters.  —  Open  and  closed  heaters  have 
their  respective  advantages  and  a  careful  study  of  the  various  influ- 
encing conditions  is  necessary  for  an  intelligent  choice.  The  follow- 
ing parallel  comparison  brings  out  a  few  of  the  distinguishing  features: 

OPEN  HEATER.  CLOSED  HEATER. 

Efficiency. 

With  sufficient  exhaust  steam  for  heat-  The  maximum  temperature  of  the  feed 
ing,  the  feed  water  may  reach  the  water  will  always  be  2  degrees  or  more 
same  temperature  as  the  steam»  lower  than  the  temperature  of  the 

Scale  and  oil  do  not  affect  the  heat        steam. 

transmission.  Scale  and  oil  deposit  on  the  tubes  and 

the  heat  transmission  is  lowered. 

Pressures. 

It  is  not  ordinarily  subjected  to  much  The  water  pressure  is  slightly  greater 
more  than  atmospheric  pressure.  than  that  in  the  boiler  when  placed 

on  the  pressure  side  of  the  pump  as 
is  customary. 

Safety. 

Sticking  of  the  back  pressure  valve  may    It  will  safely  withstand  any  pressure 
cause  it  to  "  blow  up  "  if  provision  is        likely  to  occur, 
not  made  for  such  an  emergency. 

Purification. 

Since  the  exhaust  steam  and  feed  water    Oil  does  not  come  in  contact  with  the 

mingle,  provision  must  be  made  for        feed  water. 

removing  the  oil  from  the  steam.  Scale  is  removed  with  difficulty. 

Scale  and  other  impurities  precipitated 

in  the  heater  are  readily  removed. 

Location. 

Must  always  be  placed  above  the  pump  May  be  placed  anywhere  on  the  pres- 
suction  and  on  the  suction  side.  sure  side  of  the  pump. 

Pumps. 

With  supply  under  suction  two  pumps    One  cold-water  pump  is  necessary, 
are  necessary  and  one  must  handle 
hot  water. 

Adaptability. 

Particularly     adaptable     for     heating    All   vacuum    or   primary   heaters    are 
systems  where  it  is  desired  to  pipe        necessarily  of  this  type, 
the  "  returns  "  direct  to  heater. 


FEED-WATER  PURIFIERS  AND  HEATERS 


505 


261.  "  Through  "  Heaters.  —  Fig.  249  shows  a  typical  installation  of 
a  through  heater  in  a  non-condensing  plant. 

It  is  evident  that  all  the  steam  must  pass  through  the  heater.  Now, 
one  pound  of  exhaust  steam  in  condensing  gives  up  approximately 


FIG.  249.     Open  Heater  Connected  as  a  "  Through  "  Heater.     Non-Condensing  Plant. 

1000   B.T.U.      Hence,  if  the   initial   temperature  of   the  feed  water 
is   50   degrees  and   the   final   temperature   210,  the  engine   furnishes 

=  6.26,  say  six  times   the    quantity  necessary  for    heating 


—  ou 

the  feed  water  to  a  maximum.  Therefore  the  area  of  the  pipe 
supplying  the  heater  with  steam  need  be  but 
one-sixth  that  of  the  main  exhaust.  With 
the  heater  connected  as  in  Fig.  249  the  connec- 
tions must  necessarily  be  the  same  size  as 
the  exhaust  pipe. 

With  this  arrangement  the  heater  cannot  be 
"cut  out"  while  the  engine  is  in  operation  and 
hence  it  is  not  adapted  for  plants  working 
continuously.  For  the  purpose  of  cutting  out  a 
heater  while  the  plant  is  in  operation  a  through 
heater  may  be  by-passed  as  in  Fig.  250.  Ad- 
vantage may  be  taken  here  of  the  permissible 
FIG  250  reduction  in  the  size  of  pipes  and  fittings,  i.e., 

valves,  etc.,  at  C  and  D  need  be  but  one-half 

the  size  of  those  at  A.     This  reduction  in  size  may  prove  to  be  a  con- 

siderable item  in  large  installations. 


WG 

2i 

V    indi 


STEAM  POWER  PLANT  ENGINEERING 


262.  Induced  Heaters.  —  Fig.  251  shows  a  typical  installation  of  an 
induced  heater  in  a  non-condensing  plant  and  Fig.  252  an  induced 
primary  heater  in  a  condensing  plant. 


COLD  WATER  SUPPLY 


FIG.  251 .    Open  Heater  Connected  as  an  "  Induced  "  Heater.     Non-Condensing  Plant. 


TO    ATMOS 


ATMOSPHERIC 
RELIEF  VALVE 


FEED  WATER 


FIG.  252.     Closed  Heater  Connected  as  an  "  Induced"  Heater.    Condensing  Plant. 

In  the  arrangement  in  Fig.  251  the  number  of  fittings  is  reduced 
to  a  minimum  and  the  heater  may  be  readily  cut  out.  Since  induced 
heaters  are  apt  to  become  air-bound,  a  vapor  pipe  or  vent  is  inserted 


FEED-WATER  PURIFIERS  AND  HEATERS  507 

in   the   top   of   the  heater  as  shown.     This  pipe  varies  from  J  to  1J 
inches  in  diameter,  depending  upon  the  size  of  heater. 

Closed  Heaters:  Am.  Elecn.,  May,  1900,  p.  236,  July,  1900,  p.  354,  Oct.,  1905, 
p.  530;  Cassier's  Mag.,  Aug.,  1903,  p.  330;  Eng.  U.S.,  Jan.  1,  1906,  p.  13;  Power, 
April,  1902,  p.  11. 

263.  Live-Steam  Heaters  and  Purifiers.  —  The  function  of  a  live- 
steam  heater  and  purifier  is  primarily  that  of  purification  and  hence  it 
is  not  ordinarily  installed  unless  the  feed  water  contains  scale-forming 
elements  such  as  sulphates  of  lime  and  magnesia.  These,  as  pre- 
viously stated,  are  not  entirely  precipitated  until  a  temperature  of 
approximately  300  degrees  F.  is  reached;  hence  no  amount  of  heating 
with  exhaust  steam  at  atmospheric  pressure  will  thoroughly  purify 
feed  water  containing  these  elements. 

Fig.  253  shows  a  section  through  a  Hoppes  live-steam  purifier. 
Since  the  purifier  is  subjected  to  full  boiler  pressure,  the  shell  and 


h 


FIG.  253.    Hoppes  Live  Steam  Purifier. 

heads  are  constructed  of  steel.  Within  the  shell  are  a  number  of 
trough-shaped  pans  or  trays  placed  one  above  another  and  supported 
on  steel  angle  ways.  Steam  from  the  boiler  enters  the  chamber  at  A 
and  comes  in  contact  with  feed  water  and  condenses.  The  water  on 
entering  the  heater  at  B  is  fed  into  the  top  pan  and,  overflowing  the 
edges,  follows  the  under  side  of  the  pan  to  the  center  and  drops  into  the 
pan  below.  It  flows  over  each  successive  pan  in  the  same  manner 
until  it  reaches  the  chamber  at  the  bottom,  whence  it  gravitates  to  the 
boiler  through  pipe  C.  As  the  steam  inclosed  in  the  shell  comes  in 
contact  with  the  thin  film  of  water,  the  solids  held  in  solution  are 
separated  and  adhere  to  the  bottom  of  the  pans  in  the  same  manner 


508 


STEAM  POWER  PLANT  ENGINEERING 


that  stalactites  form  on  the  roofs  of  natural  caves.  Authentic  tests 
show  that  live  steam  heaters  do  not  increase  the  boiler  efficiency  and 
that  they  merely  act  as  purifiers.  (Power,  March  31,  1908,  p.  498.) 
The  purifier  should  be  set  in  such  a  position  as  will  bring  the  bottom 
of  the  shell  two  feet  or  more  above  the  water  level  of  the  boilers,  as  in 
Fig.  254.  N  is  the  feed  pipe  from  pump  to  purifier  and  should  be  pro- 


FIG.  254.     Typical  Installation  of  a  "  Live  Steam"  Purifier. 

vided  with  a  check  valve.  D  is  the  gravity  pipe  through  which  the 
purified  water  flows  to  the  boiler.  This  pipe  should  be  carried  below 
the  water  level  of  the  boilers  and  all  branch  pipes  should  be  taken  off 
below  the  water  line.  Pipe  L  leads  from  top  of  pipe  S  to  pump 
or  other  steam-using  device.  This  is  necessary  in  order  that  air 
and  other  non-condensable  gases  liberated  from  the  water  may  be 
removed  from  the  purifier,  which  would  otherwise  become  air-bound. 
In  the  illustration  the  feed  pump  takes  its  supply  from  an  exhaust 
steam  heater  C.  The  purifier  is  provided  with  a  suitable  by-pass  so 
that  the  water  may  be  fed  directly  to  the  boiler  when  necessary. 

Live  Steam  Heated  Feed  Water:  Elec.  Engr.,  Lond.,  June  29,  1906;  Am.  Elecn., 
May,  1900,  p.  214;  Elec.  Rev.,  Lond.,  May  20,  1898,  p.  667;  Eng.  Rec.,  Aug.  30,  1898, 
p.  467;  Power,  March  31,  1908,  p.  498. 

264.  Economizers.  —  Fig.  256  gives  a  general  view  of  a  Green  econo- 
mizer, illustrating  a  typical  flue  gas  heater.  It  consists  of  a  series  of 
cast-iron  tubes  9  to  12  feet  in  length  and  4i-inches  in  diameter,  which 


FEED-WATER  PURIFIERS  AND  HEATERS  509 


FIG.  255.     Typical  Installation  of  Primary  and  Secondary  Heaters. 


510 


STEAM  POWER  PLANT  ENGINEERING 


FEED-WATER  PURIFIERS  AND  HEATERS 


511 


are  arranged  vertically  in  sections  of  various  widths  across  the  main  flue 
between  boiler  and  chimney.  When  in  position  the  sections  are  con- 
nected by  top  and  bottom  headers,  and  the  headers  are  connected 
to  branch  pipes  running  lengthwise,  one  at  the  top  and  the  other  at 
the  bottom.  Both  of  the  branch  pipes  are  outside  the  brickwork  which 
incloses  the  apparatus.  The  waste  gases  are  led  to  the  economizer  by 
the  ordinary  flue  from  the  boiler  to  the  chimney,  but  a  by-pass  must  be 
provided  for  use  when  the  economizer  is  out  of  service  for  cleaning  or  for 
repairs.  The  feed  water  is  forced  into  the  economizer  through  the  lower 
branch  pipe  nearest  the  point  of  exit  of  gases,  and  emerges  through  the 
upper  branch  pipe  nearest  the  point  where  the  gases  enter.  Each 
tube  is  encircled  with  a  set  of  triple  overlapping  scrapers  which  travel 
continuously  up  and  down  the  tubes  at  a  slow  rate  of  speed,  the  object 
being  to  keep  the  external  surfaces  free  from  soot.  The  mechanism 
for  working  the  scrapers  is  placed  on  top  of  the  economizer,  outside 
the  chamber,  and  the  motive  power  is  supplied  either  by  a  belt  from 
some  convenient  shaft  or  small  independent  engine  or  motor.  The 
power  for  operating  the  gearing  varies  from  1  to  J  horse  power  per 
1000  square  feet  of  economizer  surface,  depending  upon  the  number 
and  length  of  tubes.  The  apparatus  is  fitted  with  blow-off  and  safety 
valves,  and  a  space  is  provided  at  the  bottom  of  the  chamber  for  the 
collection  of  soot. 

Fig.  257  shows  a  typical  installation  of  a  fuel  economizer. 


FIG.  257.     A  Typical  Economizer  Installation. 


(Eng.  Rec.) 


265.   Value  of  Economizers.  —  The  general  conclusion  drawn  from 
current  practice  is  that  an  economizer  installation  results  in  — 
A  small  annual  saving  in  cost  of  operating  the  plant. 


512  STEAM  POWER  PLANT  ENGINEERING 

Decreased  wear  and  tear  on  the  boilers  due  to  the  higher  feed-water 
temperature. 

A  large  storage  of  hot  water  for  sudden  increase  in  load. 

Purification  of  the  feed  water  due  to  the  high  temperature  in 
the  economizer.  The  scale-forming  elements  do  not  bake  hard  on  the 
economizer  tubes  as  they  would  in  the  boiler  where  the  heat  from  the 
fire  is  more  intense,  but  make  a  muddy  deposit  readily  removed  by 
blowing  off. 

266.  Factors  Determining  Installation  of  Economizers.  —  The  factors 
to  be  considered  before  installing  an  economizer  are: 

The  nature  of  the  auxiliary  machinery,  direct  connected  or  belted. 

Method  of  heating  the  feed  water;  whether  vacuum  and  atmos- 
pheric heaters  are  used  and  whether  all  or  part  of  the  auxiliary  steam 
is  used  for  heating. 

Initial  temperature  of  the  feed  water;  whether  the  feed  is  taken 
from  the  hot  well  or  from  a  cold  supply. 

Rise  in  temperature  due  to  economizer. 

Cost  of  economizer.  An  approximate  price  is  $15  per  tube  erected, 
on  a  basis  of  15  square  feet  per  tube.  The  heating  surface  is  rated  at 
3  to  5  square  feet  per  boiler  horse  power. 

Cost  of  additional  building  space. 

Reduction  in  boiler-heating  surface  made  possible  by  the  econo- 
mizer. 

Extra  cost  of  stack  or  forced-draft  apparatus  necessary  to  com- 
pensate for  loss  of  draft  due  to  economizer.  The  economizer  lessens 
the  draft  by  increasing  the  resistance  between  boilers  and  chimney 
and  by  reducing  the  chimney  temperature.  Where  the  installation  of 
an  economizer  decreases  the  normal  temperature  of  the  chimney  from 
say  550  degrees  to  350  degrees  F.,  the  reduction  in  draft  is  approxi- 
mately 25  per  cent. 

Total  cost  of  economizer  plant.  This  depends  largely  upon  the 
design  and  varies  from  $4  to  $7  per  boiler  horse  power. 

Interest,  depreciation,  repairs,  operation,  taxes,  and  insurance. 

Table  67  gives  the  results  of  economizer  tests. 

367.  Temperature  due  to  Use  of  Economizer.  —  The  rise  in  temper- 
ature of  feed  water  due  to  the  use  of  an  economizer  may  be  approxi- 
mated from  the  following  empirical  formula  advocated  by  the  Green 
Economizer  Company : 


x  = 


(120) 


FEED-WATER  PURIFIERS  AND  HEATERS 


513 


in  which 

x  =  rise  in  temperature  of  the  feed  water. 
Tl  =  temperature  of  flue  gas  entering  economizer. 
^  =  temperature  of  feed  water  entering  economizer. 
w  =  pounds  of  feed  water  per  boiler  horse  power  per  hour. 
G  =  pounds  of  flue  gas  per  pound  of  combustible. 
C  =  pounds  of  coal  per  boiler  horse  power  per  hour. 
y  =  square  feet  of  economizer  heating  surface  per  boiler  horse 
power. 

o 


FIG.  258. 


Referring  to  Fig.  258,  the  ordinates  represent  temperatures,  and 
abscissas  the  path  of  the  flue  gas  and  the  water  in  the  economizer. 
The  flue  gas  enters  the  economizer  at  c  with  temperature  Tl  and 
leaves  at  a  with  temperature  T.  The  feed  water  enters  at  6  with 
temperature  tl  and  leaves  at  d  with  temperature  ^  +  x. 

The  algebraic  mean  temperature  difference  D  between  the  flue  gas 
and  the  feed  water  will  be 


wx 


--  2) 

=  T,  -  (t,  +  x)  +  T,  -  oa  -  t, 
2 

:        .  _Ti_ti_>L±s*.  (124) 

Now,  wx  =  B.T.U.   absorbed   by  the  feed  water  per    boiler  horse 

power  and 

B.T.U.  given  up  to  the  feed  water  by  the  flue  gas  for 
each    degree    reduction  in   temperature  (S  =  specific 
heat  of  the  flue  gas);  therefore 
GCS  =  total  reduction  in  temperature  of  the  flue  gas;   that  is, 


GCS 


oa. 


(125) 


514  STEAM  POWER  PLANT  ENGINEERING 

Substituting  (125)  in  (124),  we  get 


wx 
' 


D-  r.-l.-  (126) 


in  which 

D  =  mean  temperature  difference  between  flue  gas  and  feed 

water,  degrees  F. 
Let  U  =  B.T.U.  absorbed  per  hour  per  square  foot  per  degree 

difference  in  temperature  and 

y  =  square  feet  of  economizer  surface  per  boiler  horse  power. 
Then  UDy  =  heat  absorbed  per  boiler  horse-power  hour. 
But  wx  =  heat  absorbed  per  boiler  horse-power  hour. 

Therefore  wx  =  UDy.  (128) 

Combining  (128)  and  (127), 


wx  =  U 
from  which  x  =     yUl~^0-  (130) 

W        W  +  (rCiS 

C/"1      2  GOT 

?/  varies  from  3.5  square  feet  to  5  square  feet  per  boiler  horse  power, 
and  U  from  2.25  to  3.3,  depending  upon  the  conditions  of  operation.* 

If  we  let  w  =  30,  S  =  0.2,  and  U  =  3.3,  and  substitute  these  values 
in  equation  (130),  it  assumes  the  form  given  by  the  Green  Economizer 
Company,  equation  (120). 

A  method  of  approximating  the  rise  in  temperature  where  the  final 
temperature  of  the  flue  gas  is  known,  is  to  assume  T%  degree  rise  in  the 
feed  water  for  each  degree  reduction  in  temperature  in  the  flue  gas. 
This  is  determined  on  the  basis  that  approximately  20  pounds  of  flue 
gas  are  generated  for  each  pound  of  combustible,  and  that  10  pounds 
of  water  are  evaporated  per  pound  of  combustible;  that  is,  2  pounds 
of  flue  gas  are  generated  for  each  pound  of  feed  water  delivered  to  the 
boiler.  Assuming  a  specific  heat  of  0.25  for  the  flue  gas,  this  gives 
2  X  .25  or  0.5  degree  rise  in  temperature  in  the  feed  water  for  each 
degree  reduction  in  the  flue  gas  temperature. 

Example:  Determine  the  rise  in  temperature  of  the  feed  water  in  a 
power  plant  of  1200  I.H.P.  Engines  use  20  pounds  of  steam  per 

*  For  D  =  600     U  =  3.25  For  D  =  400     U  =  2.75 

500  3.00  300  2.25 


FEED-WATER  PURIFIERS  AND  HEATERS 


515 


I.H.P.  hour;  auxiliaries  use  the  equivalent  of  12  per  cent  of  main 
engine  steam;  vacuum  25  inches;  feed-water  supply  50  degrees;  3.7 
pounds  of  coal  are  burned  per  hour  per  boiler  horse  power;  flue  gas 
temperature  550  degrees  F.;  steam  pressure  150  pounds  gauge. 

The  vacuum  and  atmospheric  heater  will  raise  the  temperature  of  the 
feed  water  from  50  to  205  degrees.  (See  preceding  problem.) 

On  the  assumption  that  20  pounds  of  flue  gas  are  generated  per 
pound  of  combustible  and  that  3.5  square  feet  of  economizer  heating 
surface  are  installed  per  boiler  horse  power,  the  notations  in  the  formula 
will  become  T1=  550,  tt=  205,  w  =  30,  G  =  20,  C  =  3.7,  y  =  3.5, 
U  =  3.3,  S  =  0.2. 

Substituting, 

=  3.5  (550-205) 

30        (5  X  30  +  20  X  3.7) 
3.3  (2  X  20  X  3.7  ) 

=  83  degrees  rise  in  temperature. 

Therefore  the  temperature  of  the  water  entering  the  boiler  will  be 
205  +  83  =  288  degrees  F. 

Economizers:  Prac.  Engr.  U.S.,  March,  1910;  Engr.  U.  S.,  March  15,  1900,  p.  69; 
Power,  July  27,  1909;  Cassier's,  March,  1900,  p.  378;  St.  Ry.  Jour.,  Oct.  31,  1903, 
1903,  p.  822. 

Economizer  Installations:  Eng.  Rec.,  March  15,  1902,  p.  247,  June  7,  1902,  p.  532, 
Nov.  1,  1902,  p.  410;  Power,  Feb.,  1905. 

Tests  of  Economizers:  Am.  Elecn.,  Nov.,  1902,  p.  518;  Elec.  Rev.,  Lond.,  Aug. 
2,  1895,  p.  148;  Eng.  Rec.,  March  9,  1901,  p.  220,  July  25,  1903,  p.  102;  Power, 
Aug.,  1897,  p.  8,  Aug.,  1904,  p.  493;  St.  Ry.  Rev.,  June,  1904,  p.  436. 

TABLE   67. 

ECONOMIZER    PERFORMANCES.* 


Temperatures. 

FIIP! 

Plants 
Tested. 

Gases 

Gases 

Water 

Water 

Gain  in 

x1  lid 

Saving. 

Entering 

Leaving 

Entering 

Leaving 

Temperature 

Economizer. 

Economizer. 

Economizer. 

Economizer. 

of  Water. 

Degrees. 

Degrees. 

Degrees. 

Degrees. 

Degrees. 

Per  Cent. 

1 

610 

340 

110 

287 

177 

16.7 

2 

505 

212 

84 

276 

192 

17.1 

3 

550 

205 

185 

305 

120 

11.7 

4 

522 

320 

155 

300 

145 

13.8 

5 

505 

320 

190 

300 

110 

10.7 

6 

465 

250 

180 

295 

115 

11.2 

7 

490 

290 

165 

280 

115 

11.0 

8 

495 

190 

155 

320 

165 

15.5 

9 

595 

299 

130 

311 

181 

16.8 

*  Transactions  American  Society  of  Mechanical  Engineers.  Vol.  XV. 


516  STEAM  POWER  PLANT  ENGINEERING 

268.  Choice  of  Feed- Water  Heating  System.  —  The  heating  of  feed 
water  and  its  delivery  to  the  boiler  in  the  most  economical  manner 
is  a  problem  involving  such  a  large  number  of  combinations  that  a 
general  analysis  is  impracticable.  The  following  discussion  of  a  spe- 
cific case  will  give  some  idea  of  the  manner  in  which  this  problem  may 
be  attacked. 

Example:  Determine  the  most  economical  manner  of  heating  the 
feed  water  for  a  power  plant  of  1000  horse  power  operating  under  the 
following  conditions:  Schedule  10  hours  per  day  and  310  days  per 
year;  load  factor  on  the  ten-hour  basis  0.8;  cost  of  coal  $2.50  per 
ton  of  2000  pounds;  heat  value  of  the  coal  13,500  B.T.U.  per  pound; 
average  boiler  efficiency  65  per  cent;  engines  use  20  pounds  of  steam 
per  I.H.P.  hour;  steam  pressure  150  pounds  absolute;  temperature  of 
cold  water  60  degrees;  vacuum  26  inches  referred  to  30-inch  barometer; 
interest  5  per  cent;  depreciation  8J  per  cent;  maintenance  1  per  cent; 
insurance  \  per  cent;  taxes  1  per  cent;  total  charges  16  per  cent; 
charges  for  attendance  and  maintenance  assumed  to  be  the  same  in 
each  case  and  credit  for  the  chimney  assumed  to  offset  debit  for  econo- 
mizer space.  Many  of  the  influencing  conditions  are  left  out  for  the 
sake  of  simplicity. 

The  most  likely  combinations  are 

(1)  Atmospheric,  all  auxiliaries  steam  driven,  water  taken  from  cold 

well. 

(2)  Same  as  1  except  that  water  is  taken  from  hot  well. 

(3)  Economizers,  auxiliaries  electrically  driven,  chimney  draft,  water 

from  cold  well. 

(4)  Vacuum  heater,  economizer,  and  electrically  driven  auxiliaries, 

fan  draft. 

(5)  Vacuum  heater,  atmospheric  heater,  and  steam  auxiliaries. 

(6)  Atmospheric  heater,  economizer,  steam  auxiliaries,  fan  draft. 

(7)  Vacuum  and  atmospheric  heaters,  economizers,  steam  auxiliaries, 

and  electrical  fan. 

(8)  Vacuum,  atmospheric  heater,  economizer,  and  chimney  draft,  aux- 

iliaries operating  condensing  except  feed  pumps  and  stoker 
engines  which  exhaust  into  the  atmospheric  heater. 

The  difference  between  the  total  heat  furnished  by  the  boiler  and 
the  heat  returned  in  the  feed  water  is  the  net  heat  put  into  the  steam 
by  the  boiler.  Evidently  the  system  which  shows  the  least  net  heat 
required  to  produce  one  horse  power  will  be  the  most  economical  as 


FEED-WATER  PURIFIERS  AND  HEATERS  517 

far  as  coal   consumption  is  concerned,  although   not   necessarily  the 
cheapest  when  both  operating  and  fixed  charges  are  considered. 

Prices  vary  so  much  that  it  is  practically  impossible  to  give  costs  of 
installations  which  will  bear  criticism  and  the  prices  taken  in  this  problem 
are  approximate  only. 

CASE   I. 

Atmospheric  heater,  auxiliaries  steam  driven,  feed  from  cold  well. 

This  arrangement  and  that  of  Case  II  are  the  most  common  in  power 
plants  of  this  size. 

The  power  consumption  of  the  auxiliaries  operating  non-condensing 
varies  from  8  to  12  per  cent  of  the  total  power  developed.  Assume 
it  to  be  10  per  cent. 

The  temperature  of  the  feed  water  leaving  the  heater  may  be 
determined  by  formula  (105). 

t0+  .98(1  +  32) 


1  +  .98 

Substituting  S  = 

.10,  X=  1146,  t, 

,  =  60, 

t       60  +  .9  X 

.10  (1146  +  32) 

1  4 

-  .9  X  .10 

=  152. 

The  net  heat  furnished  by  the  boiler  to  produce  one  indicated  horse- 
power hour  in  the  engine  is  evidently  the  heat  necessary  to  raise 
20  +  10  per  cent  of  20  =  22  pounds  of  water  from  152  degrees  F.  to 
steam  at  150  pounds  pressure;  i.e.,  the  net  heat  furnished  is 

22  X  1071.2  =  23,564  B.T.U. 

Now,  1  I.H.P.  =  2545  B.T.U. 

Therefore  the  heat  efficiency  of  this  arrangement  is 

2545 

L0.8  per  cent. 


Probable  First  Cost. 

Steam  pumps  ...............................................  $400.00 

Condenser  with  steam-driven  air  and  circulating  pumps  ..........  3000.00 

1000  horse-power  open  heater  ..................................  480.00 

Piping  ...........................................  .  .  1200.00 


$5080.00 


518  STEAM  POWER  PLANT  ENGINEERING 

Fuel  Consumption. 

Average  horse-power  hours  per  year  =  1000  (rated  horse  power)  X  0.8  (curve 
load  factor)  X  310  (days  per  year)  X  10  (hours  per  day)  =  2,480,000. 

Pounds  of  coal  per  I.H.P.  hour  =  net  heat  furnished  per  I.H.P.  hour  -f-  net  heat 
absorbed  by  the  boiler  per  pound  of  coal  =  23,564  -r-  (13,500  X  0.65)  =  2.68. 

Tons  per  year  ^80,000  X  2.68  J 
2000 

Fuel  and  Fixed  Charges. 

Fuel,  3323  tons  at  $2.50    $8,308.00 

Fixed  charges,  16  per  cent  of  $5080 812.00 


$9,120.00 
CASE  II. 

Same  as  Case  I,  except  that  feed  is  taken  from  the  hot  well.  This 
arrangement  is  possible  only  when  the  condensing  water  is  suitable 
for  feed  purposes. 

Assume  the  temperature  of  the  water  from  the  hot  well  as  it  enters 
the  heater  to  be  110  degrees. 

The  temperature  of  the  feed  water  leaving  the  heater  will  then  be 
198  degrees  (from  formula  (105)). 

Net  heat  furnished  =  22  X  1025.2  =  22,554  B.T.U. 

OK4K 

Efficiency  =    *  JT   =11.3  per  cent. 
.2^,554 


.  Pounds  of  coal  per  I.H.P.  =  =  2.62. 

-t)0 


Tons  per  year-          3248. 


Fuel  and  Fixed  Charges. 

Fuel,  3248  tons  at  $2.50  ....................................  $8,120.00 

Fixed  charges  (same  as  Case  I)  ..............................      812.00 


$8,932.00 
CASE  III. 

Economizers,  auxiliaries  electrically  driven,  chimney  draft,  water 
from  the  cold  well. 

Practice  gives  an  average  of  3  per  cent  of  the  main  engine  output  as 
the  power  required  to  operate  the  electrical  auxiliaries  in  a  plant  of  this 
size. 


FEED-WATER  PURIFIERS  AND  HEATERS  519 

The  temperature  of  the  feed  water  leaving  the  economizer  may  be 
determined  from  formula  (120). 


Substituting, 

3.5(550-60) 


2  X  20  X  3.7 

Temperature  of  feed  water  entering  heater  =119  +  60  =  179  degrees. 
Net  heat  furnished  =  (20  +  3  per  cent  of  20)  X  1044.2  =21,510  B.T.U. 


Efficiency  =  -  =  11.8  per  cent. 

21,510 

Probable  First  Cost. 
Economizers  ____  ..........................................  $3,500.00 

Motor  feed  pump  ..........................................     600.00 

Condenser  with  electrically  driven  air  and  circulating  pump  .....  6,000.00 

Piping  and  wiring  .......  .  .................................   1,000.00 


$11,100.00 

Fuel  Consumption. 

Pounds  of  coal  per  I.H.P.  hour  =  — ^i5 2.45. 

13,500  X  .65 

2,480,000  X  2.45 

Tons  per  year= =  3038. 

2000 

Fuel  and  Fixed  Charges. 

Fuel,  3038  tons  at  $2.50 $7,595.00 

Fixed  charges,  16  per  cent  on  $11,100    1,776.00 


$9,371.00 
CASE  IV. 

Vacuum  heater,  economizer,  electrically  driven  auxiliaries,  fan  draft. 

The  vacuum  heater  may  be  relied  upon  to  raise  the  temperature  of 
the  feed  water  to  110  degrees. 

The  economizer  will  increase  this  107  degrees  (from  formula  (120)), 
giving  the  feed  water  a  temperature  of  217  degrees  as  it  enters  the 
heater. 

The  electrical  fan  for  the  mechanical-draft  system  will  require  approx- 
imately 2  per  cent  of  the  main  system  engine  power,  making  a  total  of 
3  +  2  =  5  per  cent  for  all  auxiliaries. 


520 


STEAM   POWER  PLANT  ENGINEERING 


Net  heat  furnished  =  (20  +  5  per  cent  of  20)  X  1006.2. 
=  21,130  B.T.U. 


Efficiency 


-  12.05  per  cent. 


Probable  First  Cost. 

For  the  sake  of  simplicity  it  is  assumed  that  the  high  first  cost  of  the 
chimney  plus  its  low  depreciation  and  maintenance  will  offset  the  low 
first  cost  of  the  mechanical-draft  system  plus  its  higher  maintenance 
and  depreciation  charges. 

Economizers  ..............................................  $3,500.00 

Motor  feed  pump  ..........................................      600.00 

Motor-driven  pumps  and  condenser  ..........................   6,000.00 

Motor-driven  fan  ..........................................      750.00 

Piping  and  wiring  ...............  ...........................    1,200.00 

Vacuum  heater  ............................................      200.00 


$12,250.00 


Fuel  Consumption. 


Pounds  of  coal  per  I.H.P.  hour  = 


13,500  X  .65 


=  2.41. 


2,480,000X2.41 

Tons  per  year  =  —  -  -  -  =  2988. 
2000 

Fuel  and  Fixed  Charges. 

Fuel,  2988  tons  at  $2.50  ....................................  $7,470.00 

Fixed  charges,  16  per  cent  of  $12,250  ........................   1,960.00 

$9,430.00 
In  like  manner  Cases  V,  VI,  VII  and  VIII  have  been  treated  and  are 


tabulated  in  the  summaries. 


SUMMARY   (1). 


Case. 

Temperature 
of  Feed 

Water. 

Power 
Consumed  by 
Auxiliaries. 

Efficiency. 

First 
Cost. 

Fuel  Cost 
per  Year. 

Cost  of 
Operation 
per  Year. 

I 

Degrees  F. 
152 

Per  Cent. 
10 

Per  Cent. 
10.8 

$5,080 

$8,308 

$9,120 

II  
Ill      . 

198 
179 

10 
3 

11.3 
11.8 

5,080 
11,100 

8,120 
7,595 

8,932 
9,371 

IV         

217 

5 

12.05 

12,250 

7,470 

9,430 

v  

208 

10 

11.4 

5,280 

7,900 

8,744 

VI  

294 

14 

12 

9,000 

7,750 

9,190 

VII  

290 

10 

12.2 

9,300 

7,380 

9,570 

VIII  

270 

8 

12.3 

8,250 

7,075 

8,395 

FEED-WATER  PURIFIERS  AND  HEATERS 
SUMMARY   (2). 


521 


Case. 

Efficiency. 

First  Cost. 

Fuel. 

Cost  per  Year. 

I 

8 

1 

g 

4 

II  
Ill  
IV  
V.  . 

7 
6 
3 
5 

1 
6 
7 
2 

7 
4 
3 
6 

2 
6 
7 
3 

VI  

4 

4 

5 

5 

VII 

2 

5 

2 

8 

VIII  

1 

3 

1 

1 

Summary  (2)  gives  the  ranking;  thus:  Case  I  is  eighth  in  point  of 
efficiency;  first  in  cheapness  of  installation;  eighth  in  yearly  cost  of 
fuel;  and  fourth  in  yearly  cost  of  operation.  Case  VIII  is  apparently 
the  best  arrangement  for  the  given  conditions. 


CHAPTER  XIII. 

PUMPS. 

269.  Classification.  —  Pumps  used  in  connection  with  steam  power 
plants  may  be  conveniently  classified  under  five  groups  according  to 
the  principles  of  action. 

1.  Piston  pumps,  in  which   motion  and  pressure  are  imparted  to 
the  fluid  by  a  reciprocating  piston,  plunger,  or  bucket.     The  action  is 
positive  and  a  certain  definite  amount  of  fluid  is  handled  per  stroke 
under  predetermined  conditions  of  pressure  and  velocity. 

2.  Centrifugal  pumps,  in  which  the  fluid  is  given  initial  velocity  and 
pressure  by  a  rotating  impeller.     The  action  is  not  positive,  as  the 
amount  of  fluid  discharged  is  not  necessarily  proportional  to  the  impeller 
displacement. 

3.  Rotary  pumps,  in  which   motion  and  pressure  are  imparted  to 
the  fluid  by  a  rotating  impeller.    The  volume  discharged  is  practically 
equal  to  the  impeller  displacement  regardless  of  pressure. 

4.  Jet  pumps,  in  which  velocity  and  pressure  are  imparted  to  the 
fluid  by  the  momentum  of  a  jet  of  similar  or  other  fluid.     The  ordinary 
steam  injector  is  the  best  known  of  this  group. 

5.  Direct-pressure  pumps,  in  which  the  pressure   of  one  fluid  acts 
directly  on  the  surface  of  another  fluid,  thereby  imparting  all  or  part 
of  its  energy  to  the  latter.     The  pulsometer  is  an  example  of  this  type. 

These  groups  may  be  variously  subdivided  as  follows: 


Piston 

Direct-acting.  . 

Fly-wheel  
Power  driven  .  . 

Volute 

j  Simplex  ... 
1  Duplex.  .  .  . 
1  Simplex.  .  . 
Duplex  
Triplex  
Single  stage 
Multi-stage 
(  Forcing.  .  . 
I  Lifting  .... 
j  Positive  .  .  . 
\  Automatic 
.  .  Lifting  .... 
.  .Lifting 

Air. 
Vacuum. 
Forcing. 
Lifting. 

Forcing. 
Lifting. 

Centrifugal 
Rotary  • 
Jet  

.  Turbine  
Power  driven  .  . 

Injector  

Pulsometer  .... 
Air-lift  

Direct  pressure 

Piston  or  plunger  pumps  are  the  most  common  in  use.  Boiler- 
feed  pumps,  city  waterworks  pumps,  and  force  pumps  are  ordi- 
narily of  this  type.  In  the  direct-acting  type,  Fig.  260,  the  water 

522 


PUMPS  523 

plunger  and  steam  piston  are  secured  to  a  single  piston  rod  and  the 
steam  pressure  is  transmitted  directly  to  the  water.  There  is  no  fly 
wheel,  connecting  rod,  or  crank.  The  velocity  of  the  delivery  is  pro- 
portional to  the  resistance  offered  by  the  water;  when  the  resistance 
equals  the  forward  effort  of  the  steam  pressure  the  pump  stops.  This 
class  of  pump  is  well  adapted  for  boiler-feeding  purposes,  since  it 
may  be  operated  as  slowly  as  suits  the  requirements  of  feeding  by 
simply  throttling  the  discharge.  The  steam  consumption  is  very 
large  in  proportion  to  the  work  performed,  since  the  steam  is  not  used 
expansively. 

Fly-wheel  pumps,  Figs.  273,  308,  are  ordinarily  classified  as  pumping 
engines.  In  this  class  steam  may  be  used  expansively,  as  sufficient 
energy  is  stored  in  a  fly  wheel  to  permit  the  drop  in  steam  pressure 
during  expansion.  These  pumps  find  wide  application  in  city  water- 
works, elevator  plants,  and  the  like,  where  high  duty  is  required.  They 
are  little  used  as  stationary  boiler  feeders,  but  are  used  to  some 
extent  in  river  boat  practice  and  in  plants  operating  continuously  for 
long  periods  at  comparatively  steady  loads.  Practically  all  sizes  of 
dry-air  pumps  and  a  number  of  large  jet  condenser  pumps  are  of  this 
type. 

Piston  pumps,  Fig.  279,  driven  by  gearing  or  belting  are  ordinarily 
classified  as  power-driven  pumps.  The  driving  power  may  be  steam 
engine,  electric  motor,  or  gas  engine.  The  single-cylinder  machine  is 
often  designated  as  a  "  simplex  "  power-driven  pump,  the  two-cylinder 
as  a  "  duplex,"  the  three-cylinder  as  a  "  triplex,"  and  so  on. 

Centrifugal  pumps,  Fig.  292,  are  supplanting  to  a  considerable  extent 
the  present  type  of  piston  pump  for  many  uses.  Though  particularly 
adapted  for  low  heads  and  large  volumes  they  are  used  in  many 
situations  requiring  extremely  high  heads.  They  are  not  as  efficient  as 
high-grade  pumping  engines,  but  the  extremely  low  first  cost  fre- 
quently offsets  this  disadvantage,  and  they  are  much  used  in  connection 
with  dry  docks,  irrigating  plants,  sewage  systems,  and  as  circulating 
pumps  in  condensing  plants. 

Rotary  pumps,  Fig.  305,  are  employed  to  a  limited  extent  in  the 
same  field  as  the  centrifugal  pump.  Being  positive  in  action,  they 
permit  of  a  much  lower  rotative  speed  for  the  same  delivery  pressure. 

Jet  pumps,  Fig.  282,  are  seldom  used  as  pumps  in  the  ordinary 
sense  of  the  word,  on  account  of  their  extremely  low  efficiency, 
but  are  frequently  employed  for  discharging  water  from  sumps. 
Their  greatest  field  of  application  lies  in  boiler  feeding  and  in  this 
respect  their  efficiency  is  comparable  with  that  of  the  average  piston 
pump. 


524 


STEAM  POWER  PLANT  ENGINEERING 


Direct-pressure  pumps  operated  by  steam,  such  as  the  "  pulsometer," 
Fig.  309,  are  used  principally  for  pumping  out  sumps,  surface  drains, 
and  the  like,  where  the  operation  is  intermittent.  Direct-pressure 
pumps  of  the  air-lift  type,  Fig.  310,  are  quite  common  and  are  used  a 
great  deal  in  situations  where  water  is  to  be  pumped  from  a  number 
of  scattered  wells. 

270.  Boiler-Feed  Pumps,  Direct-Acting  Duplex.  —  Figs.  259  and  260 
illustrate  a  typical  duplex  boiler-feed  pump,  which  consists  virtually  of 


AIR 
CHAMBER 


•TEAM    SUPPLY 


FIG.  259.     Typical  Duplex  Pump. 

two  direct-acting  pumps  mounted  side  by  side,  the  water  ends  and  the 
steam  ends  working  in  parallel  between  inlet  and  exhaust  pipe.  The 
piston  rod  of  one  pump  operates  the  steam  valve  of  the  other  through 
the  medium  of  bell  cranks  and  rocker  arms.  The  pistons  move  alter- 
nately, and  one  or  the  other  is  always  in  motion,  the  flow  of  water 
being  practically  continuous. 

In  general  construction  the  steam  pistons  and  valves  are  similar 
to  those  of  steam  engines.  The  valves  in  duplex  pumps,  however, 
have  no  lap.  In  order  to  reduce  the  valve  travel  to  a  minimum, 
and  still  have  sufficient  bearing  surface  between  the  steam  ports  and 
the  main  exhaust  ports  to  prevent  the  leakage  of  steam  from  one  to 
the  other,  separate  exhaust  ports  are  provided  which  enter  the  cylinder 
at  nearly  the  same  point  as  the  steam  ports.  This  arrangement  offers 


PUMPS 


525 


a  simple  means  of  cushioning  the  piston  by  exhaust  steam,  thus  pre- 
venting it  from  striking  the  cylinder  heads  at  the  ends  of  the  stroke. 
The  valves  of  the  duplex  pump  having  no  lap  would,  if  connected 
rigidly  to  the  valve  stem,  open  one  port  as  soon  as  the  other  had  been 
closed,  at  about  mid-stroke  of  the  piston,  thus  cutting  down  the  stroke 


DISCHARGE 


FIG.  260.     Section  Through  a  Typical  Duplex  Boiler-Feed  Pump. 

to  about  one-fourth  the  usual  length.  To  obviate  this  difficulty  the 
valves  are  given  considerable  lost  motion  by  allowing  sufficient  clear- 
ance between  the  lock  nuts  on  the  valve  stem;  the  latter,  therefore, 
imparts  no  motion  to  the  valve  until  the  piston  operating  it  has  nearly 
completed  the  stroke.  The  lost  motion  between  valves  and  lock 
nuts  renders  it  impossible 
to  stop  the  pump  in  any 
position  from  which  it  can-  v*tvc  STEM 
not  be  started  by  simply 
admitting  steam,  and 
therefore  the  pump  has 
no  dead  centers.  When 

one  piston  moves  to   the  1 IIOJU  I  PISTON  ROD 

end  of  the  stroke  it  pulls 

or    pushes    the    opposite 

valve  to  the  end  of  its  travel;  then  when  the  piston  starts  back  to  the 

other  end  of  its  stroke  the  valve  remains  stationary,  owing  to  the  lost 

motion,  until   the  piston  has   completed  about  one-half  the  stroke. 


FIG.  261. 


526 


STEAM  POWER  PLANT  ENGINEERING 


PISTON   ROD 


During  this  time  the  opposite  piston  has  completed  a  full  stroke  and 
the  valve  operated  by  it  will  have  opened  the  steam  port   wide,  so 

that  while  one  valve  covers 
both  steam  ports  the  other 
is  at  the  end  of  its  travel. 
In  some  makes  of  pumps 
the  stem  is  rigidly  attached 
to  the  valves,  the  lost  mo- 
tion being  adjusted  outside 
the  steam  chest  as  shown 
in  Figs.  261  and  262,  which 
represent  two  common  constructions  of  duplex  valve  gear. 

Fig.  263  shows  the  valve  and  piston  in  the  position  occupied  at  the 
commencement  of  the  stroke. 
At  one  end  of  the  valve  the 
steam  port  P  is  open  wide  and 
at  the  opposite  end  the  exhaust 
port  E  is  open  wide.  When  the 
piston  nears  the  opposite  end  of 
the  stroke  and  reaches  the  posi- 
tion shown  in  Fig.  264  the  steam 
escape  through  the  exhaust  port 
E  is  cut  off  by  the  piston,  and 
since  the  steam  port  is  closed,  the 
remaining  steam  is  compressed 
between  the  piston  and  cylinder  head,  thus  arresting  the  motion  of 
the  piston  gradually  without  shock  or  jar. 

The  construction  of  the  water 
end  of  single-cylinder  and  du- 
plex pumps  is  practically  the 
same;  any  slight  differences 
which  may  be  found  are  con- 
fined to  minor  details  which  in 
no  way  affect  the  general  design 
or  operation  of  the  pump. 
The  piston  is  double  acting, 
the  single-acting  cylinder  being 
confined  to  power  pumps  or  to 
steam  pumps  intended  for  very 
high  pressures.  In  the  old-style  pumps  it  was  the  custom  to  use  one 
large  valve  with  a  lift  sufficient  to  give  the  required  passage,  but  in 
modern  practice  the  required  area  is  divided  among  several  small 


FIG. 263. 


FIG.  264. 


PUMPS 


527 


valves,  so  that  each  one  is  easily  and  cheaply  removed  in  case  of 
accident  or  wear,  and  slip  is  lessened.* 

The  valves  are  carried  by  two  plates  or  decks,  the  suction  valves 
being  attached  to  the  lower  plate  and 
the  delivery  valves  to  the  upper  one, 
as  shown  in  Fig.  260. 

•The  valves  in  practically  all  boiler- 
feed  pumps  are  of  the  flat  disk  type, 
Fig.  265,  held  firmly  to  the  seat  by 
conical  springs  and  guided  by  a  bolt 
through  the  center. 

All  pumps  are  provided  with  an  air 
chamber  on  the  discharge  side,  which 
acts  as  a  cushion  for  the  water,  pre- 

v  j   .  FIG.  265.   A  Typical  Pump  Disk- Valve. 

vents  excessive  pounding,  and  insures 

a  uniform  flow.     Fig.  266  shows  a  section  through  the  steam  end  of 

a  compound  duplex  pump. 


FIG.  266.     Section  Through  Steam  Cylinders  of  a  Typical  Compound  Duplex  Pump. 

271.  Feed  Pumps  with  Steam-Actuated  Valves.  —  Single-cylinder 
direct-acting  pumps,  Fig.  267,  are  ordinarily  operated  by  steam-actu- 
ated valves.  The  steam  enters  the  chest  C  and  passes  to  the  left  through 
the  annular  opening  A  formed  between  the  reduced  neck  of  the  valve 
and  the  bore  of  the  steam  chest.  It  is  thus  projected  against  the 
inside  surface  of  the  valve  head  H  before  escaping  through  the  port  P 
and  passing  to  the  cylinder.  Both  the  pressure  and  impulse  due  to 
velocity  acting  on  the  valve  head  H  tend  to  close  or  restrict  the 

*  The  modern  Riedler  pump  is  an  exception.  See  Engineer,  U.S.,  Nov.  15,  1907, 
p.  1040. 


528 


STEAM   POWER  PLANT  ENGINEERING 


admission  port  by  forcing  the  valve  to  the  left.  On  reaching  the 
cylinder  and  forcing  the  piston  X  toward  the  right,  the  pressure  of 
the  steam  upon  the  opposite  side  of  the  valve  head  H  is  pressing  the 
valve  to  the  right,  a  movement  which  would  give  the  admission  more 
port  opening  at  A  and  deliver  more  steam  to  the  cylinder.  The  valve 
then  holds  a  position  depending  upon  the  relative  intensity  of  the  two 
pressures,  which  tend  to  move  it  in  opposite  directions,  the  admis- 
sion steam,  tending  to  close  the  valve,  and  cylinder  steam,  tending 


FIG.  267.     Marsh  Boiler-Feed  Pump.     A  Typical  Steam-Actuated  Valve  Gear. 

to  open  the  valve  wider.  The  steam  valve,  therefore,  is  always  in  a 
balanced  position.  The  steam  piston  is  grooved  at  the  center,  form- 
ing a  reservoir  for  live  steam  R  which  is  supplied  from  the  upper 
chamber  of  the  steam  chest  by  passage  E  to  the  cylinder  cap  S, 
and  thence  by  tube  M  and  the  hollow  piston  rod  V.  The  steam 
in  this  annular  piston  space  reverses  the  steam  valve  by  pressing 
alternately  against  the  outer  surfaces  of  the  valve  heads  H  through  the 
connecting  passages  0,  0  near  each  end  of  the  cylinder.  The  tappets  T 
are  for  the  purpose  of  moving  the  valve  by  hand  in  case  it  fails  to 


PUMPS 


529 


move  automatically.  Steam-actuated  valves  are  not  as  positive  in 
action  as  mechanically  operated  valves,  and  hence  are  little  used  in 
situations  where  positive  action  is  essential,  as  in  fire-pump  service. 

273.  Air  and  Vacuum  Chambers.  —  Air  chambers  in  piston  pumps 
are  for  the  purpose  of  causing  a  steady  discharge  of  water  and  of 
reducing  excessive  pounding  at  high 
speeds  by  providing  a  cushion  for  the 
water.  The  water  discharged  under  pres- 
sure compresses  the  air  in  the  air  cham- 
ber somewhat  above  the  normal  pressure 
of  discharge  during  each  stroke  of  the 
water  piston,  and  when  the  piston  stops 
momentarily  at  the  end  of  the  stroke 
the  air  expands  to  a  certain  extent  and 
tends  to  produce  a  uniform  rate  of  flow. 

The  volume  of  the  air  chamber  varies 
from  2  to  3J  times  the  volume  of  the 
water  piston  displacement  in  single- 
cylinder  pumps,  and  from  1  to  2J  times 
in  the  duplex  type.  High-speed  pumps  are  provided  with  air  cham- 
bers of  from  5  to  6  times  the  piston  displacement.  The  water  level 
in  the  air  chamber  should  be  kept  down  to  one-fourth  the  height  of 
the  chamber.  In  slow-running  pumps  sufficient  air  may  be  carried 
into  the  pump  chamber  along  with  the  water,  but  with  high  speeds  a 
large  part  of  the  air  will  be  discharged,  and  air  must  be  forced  into  the 
chamber  by  mechanical  means.  The  larger  the  chamber  the  more 
uniform  will  be  the  discharge  pressure. 


FIG.  268.     Forms  of  Vacuum 
Chambers. 


(A) 


(B) 


(C) 


FIG.  269.     Different  Arrangements  of  Vacuum  Chambers. 


Vacuum  chambers  are  frequently  provided  for  the  purpose  of  main- 
taining a  uniform  flow  of  water  in  the  suction  pipe  and  assisting  in  the 
reduction  of  slip.  Such  chambers  should  be  of  slightly  greater  volume 


530 


STEAM  POWER  PLANT  ENGINEERING 


than  the  suction  pipe  and  of  considerable  length  rather  than  diameter. 
Fig.  268  illustrates  two  designs  commonly  used.  The  one  in  Fig.  268  (B) 
should  be  placed  in  such  a  position  as  to  receive  the  impact  of  the 
column  of  water  in  the  suction  pipe  as  illustrated  in  Figs.  269  (A), 
269  (B),  and  269  (C).  The  chamber  illustrated  in  Fig.  268  (A)  should 
be  placed  in  the  suction  pipe  below  but  close  to  the  pump. 

273.   Water  Pistons  and  Plungers.  —  In  cold-water  pumps  the  water 
pistons  are  usually  packed  with  some  kind  of  soft  packing.     Fig.  270  (A) 


G       a 


R          R 


(A)  (B)  (C) 

FIG.  270.    Types  of  Water  Pistons. 

shows  the  details  of  a  piston  with  square  hydraulic  packing.  The  body  E 
is  fastened  to  the  piston  rod  by  nut  (7;  packing  is  placed  at  D,  and 
follower  F  is  forced  up  by  the  nut  B  and  locked  by  nut  A.  For  large 


FIG.  271.     Plunger  with  Metal  Packing  Ring. 

sizes  the  design  is  the  same  except  that  the  follower  is  set  up  by  a 
number  of  nuts  near  the  edge.  In  hot-water  pumps  the  pistons  are 
often  packed  by  means  of  metallic  piston  rings  R,  R,  Fig.  270  (C),  similar 
to  those  in  steam  pistons,  or  merely  by  water  grooves  G,  G,  Fig.  270  (B). 


PUMPS 


531 


The  water  end  is  often  fitted  with  a  plunger  instead  of  a  piston,  as  in 
Figs.  271  to  273.  The  piston  is  more  compact,  but  the  plungers  do 
not  require  a  bored  cylinder,  so  that  the  first  cost  is  not  materially 
different. 

Fig.  271  shows  a  plunger  with  metal  packing  ring.  When  leakage 
becomes  excessive  it  is  necessary  to  renew  the  ring,  which  is  readily 
removed. 

In  Fig.  272  the  plunger  is  packed  with  hydraulic  packing  as  in  the 
follower  type  of  pump  piston.  The  great  difficulty  with  the  above 


Pro.  272.     Plunger  with  Hydraulic  Packing. 

types  of  piston  and  plunger  is  in  keeping  the  packing  tight  or  in  know- 
ing when  it  is  leaking,  and  the  trouble  necessary  to  replace  the  packing. 
The  outside  packed  plunger,  Fig.  273,  obviates  these  disadvantages  to  a 
great  extent,  since  leakage  is  readily  detected  and  repacking  is  performed 
without  removing  the  cylinder  heads.  In  dirty,  dusty  locations,  how- 
ever, the  piston  pump  or  inside  packed  plunger  is  to  be  preferred, 
since  the  abrasive  action  of  the  dust  renders  outside  packing  difficult. 
Fig.  273  illustrates  a  high-duty  elevator  pump  with  outside  packed 
plunger. 

274.  Performance  of  Piston  Pumps.  —  Direct-acting  pumps  as  a 
class  are  wasteful  of  fuel  and  low  in  efficiency,  due  largely  to  the  non- 
expansive  use  of  steam.  The  average  small  duplex  boiler-feed  pump 
uses  from  100  to  200  pounds  of  steam  per  I.H.P.  hour,  depending 
upon  the  speed,  and  the  mechanical  efficiency  varies  from  50  per  cent 


532 


STEAM  POWER  PLANT  ENGINEERING 


PUMPS 


533 


to  90  per  cent.  When  new  and  in  proper  working  condition  the 
mechanical  efficiency  is  seldom  less  than  85  per  cent;  but  such  pumps, 
as  a  rule,  are  given  scant  attention,  and  the  average  efficiency  is  not 
far  from  65  per  cent.  The  term  "  mechanical  efficiency "  in  this 
connection  refers  to  the  ratio  of  the  actual  water  horse  power  to 
the  indicated  horse  power  of  the  steam  cylinder.  The  loss  includes  the 
slip  of  the  piston  and  valves.  A  steam  consumption  of  150  pounds  per 
I.H. P.  hour  with  mechanical  efficiency  of  65  per  cent  is  equivalent  to 
a  power  consumption  of  about  5  per  cent  of  the  rated  boiler  capacity, 
although  if  the  exhaust  steam  is  used  for  feed-water  heating  the  actual 
heat  consumption  may  be  but  1  to  1.5  per  cent.  Compound  direct- 
acting  pumps  running  non-condensing  use  from  50  to  100  pounds  of 
steam  per  I.H. P.  hour.  Single-cylinder  fly-wheel  pumps  of  the  slow- 
speed  type,  running  non-condensing,,  use  about  50  pounds  of  steam 
per  I.H. P.  hour.  Multi-cylinder  fly-wheel  pumps  of  the  high-duty 
type  use  about  25  pounds  per  I.H. P.  hour  when  running  non-condens- 
ing, and  as  low  as  10  pounds  when  operating  condensing.  High-grade 
direct-connected  motor-driven  power  pumps  have  a  mechanical  efficiency 
from  line  to  water  load,  at  normal  rating,  of  about  80  per  cent.  The 
efficiency  of  geared  pumps  at  normal  rating  varies  with  the  character  of 
the  gearing  and  the  degree  of  speed  reduction,  and  may  range  anywhere 
from  40  to  70  per  cent. 


Steam  Consumptlon,Eb.  Per  I.H.  P.Hr. 

i  1  1  I 

Effect  of  Speed 
on  the 
Economy  of  Small  Direct-  Acting 
Steam  Pumps 

\l 

A     16± 
B     12  x 

10  x  12 

7)^x12 

Duplex 
Simplex 

\ 

\B 

\ 

\^ 

"  —  *-     *L. 

• 
- 

—  • 

25                50                75               100               125              150               175              2<X 

Number  of  Single  Strokes  .Pet  Minute 
FIG.  274. 

The  steam  consumption  of  all  direct-acting  boiler  pumps  decreases 
with  the  increase  in  speed.  This  is  illustrated  by  curve  B,  Fig.  274, 
plotted  from  the  tests  of  a  12  x  7J  x  12  direct-acting  single-cylinder 
pump  at  Armour  Institute  of  Technology,  and  curve  A  based  on  experi- 


534 


STEAM  POWER  PLANT  ENGINEERING 


ments  with  a  16  x  12  duplex  fire  pump  at  Massachusetts  Institute  of 
Technology. 

Fig.  275  gives  the  details  of  the  performance  of  a  12  x  7J  x  12  Marsh 
boiler-feed  pump  at  the  Armour  Institute  of  Technology. 


7000 


Varying  Speed 
Size  of  Pump-  12"x  7^'x  12" 
216  Gal.  per  Min.  at  100  Strokes 


40          50          60          70 
Single  Strokes  per  Min. 

FIG.  275. 


90         100 


The  determination  of  the  power  consumption  of  a  boiler-feed  pump 
is  best  illustrated  by  the  following  example. 

Example:  A  small  direct-acting  duplex  pump  uses  150  pounds  of 
steam  per  I.H.P.  hour.  Gauge  pressure  150  pounds  per  square  inch; 
feed-water  temperature  64  degrees  F.  Required  the  per  cent  of  rated 
boiler  capacity  necessary  to  operate  the  pump. 

The  head  pumped  against,  150  pounds  per  square  inch,  is  equivalent 
to  150  X  2.3  =  345  feet  of  water. 


PUMPS  535 

The  friction  through  the  valves,  fittings,  and  pipe,  and  the  vertical 
distance  between  suction  and  feed-water  inlet,  are  assumed  to  be  equiva- 
lent to  20  per  cent  of  the  boiler  pressure,  giving  a  total  head  of  150  + 
30  =  180  pounds  per  square  inch,  or  414  feet  of  water. 

A  boiler  horse  power,  taking  into  consideration  leakage  losses  and 
the  steam  used  by  the  feed  pump,  will  be  equivalent  to  the  evapora- 
tion of  approximately  32  pounds  of  water  per  hour  from  a  feed  tem- 
perature of  64  degrees  F.  to  steam  at  150  pounds  gauge. 

The  actual  work  done  in  pumping  32  pounds  of  water  against  a  head 
of  414  feet  is 

414  X  32  =  13,248  foot-pounds. 

This  corresponds  to 

er> 


60  X  33,000 

The  total  heat  of  one  pound  of  steam  above  64  degrees  F.  is  1161.2 
B.T.U.  The  heat  delivered  to  the  pump  per  I.H.P.  hour  is 

1161.2  X  150  =  174,180  B.T.U. 

The  amount  used  by  the  pump  for  each  boiler  horse  power,  disregard- 
ing efficiency,  is 

174,180  X  0.0067  =  1167  B.T.U.  per  hour. 

The  mechanical  efficiency  of  the  average  feed  pump  ranges  from  50 
to  85  per  cent,  depending  upon  its  condition  and  the  number  of  strokes 
per  minute.  Assuming  it  to  be  65  per  cent,  the  heat  used  by  the  pump 
per  hour  to  deliver  32  pounds  of  water  into  the  boiler  is 

1167  -4-  0.65  =  1795  B.T.U. 

A  boiler  horse  power  is  equivalent  to  33,320  B.T.U.  per  hour.  There- 
fore the  per  cent  of  boiler  output  necessary  to  operate  the  pump  is 


If  the  exhaust  steam  is  used  for  heating  the  feed  water,  the  steam  con- 
sumption will  be  1.37  per  cent  of  the  boiler  capacity,  thus:  The  weight 
of  steam  consumed  per  boiler  horse-power  hour 

1.54  pounds.  ;     ;^, 

Allowing  10  per  cent  for  condensation,  the  heat  in  the  exhaust  avail- 
able for  heating  the  feed  water  is 

X  0.90  X  1.54  =  1340  B.T.U.* 
*  Surface  Condenser  Plant. 


536  STEAM  POWER  PLANT  ENGINEERING 

1795  —  1340  =  455  B.T.U.,  or  the  net  heat  required  by  the  pump 
per  hour  to  deliver  32  pounds  of  water  to  the  boiler. 
The  per  cent  of  boiler  output  necessary  to  operate  the  pump  is 


Pump  performances  are  generally  given  in  terms  of  the  foot-pounds 
of  work  done  by  the  water  piston  per  thousand  pounds  of  dry  steam  or 
per  million  B.T.U.  consumed  by  the  engine,  thus: 
i       nn^r  _  Foot-pounds  of  work  done       innn 

1.        l^Uty   —  .  -      -  -  —  A    1UUU. 

Weight  of  dry  steam  used 

2.     Duty  =  -    Foot-pounds  of  workdone  -    x  1 
Total  number  of  heat  units  consumed 

(See  A.S.M.E.  code  for  conducting  duty  trials  of  pumping  engines, 
Trans.  A.S.M.E.,  12-530,  563.) 

Example:  A  compound  feed  pump  uses  100  pounds  of  steam  per 
I.H.P.  hour;  indicated  horse  power,  48;  capacity,  400  gallons  per 
minute;  temperature  of  water,  200  degrees  F.;  total  head  pumped 
against,  175  pounds  per  square  inch;  steam  pressure,  100  pounds  gauge; 
moisture  in  the  steam,  3  per  cent.  Required  the  duty  on  the  dry  steam 
and  on  the  heat-unit  basis. 

175  pounds  per  square  inch  is  equivalent  to  175  X  2.4  =  420  feet  of 
water  at  200  degrees  F. 

Weight  of  400  gallons  of  water  at  200  degrees  F.  =  400  X  8.03  = 
3212  pounds. 

Work  done  per  minute  =  3212  X  420  =  1,349.040  foot-pounds. 

Weight  of  dry  steam  supplied  per  minute  =  10Q  *  48   X  0.97  =  77.6 

pounds. 

B.T.U.  supplied  per  minute  =  10°  X  48   (0.97  X  876.2  +  308.8-200 

ou 

4-  32)  =  79,256. 
Duty  per  thousand  pounds  of  dry  steam 

=  1?349?040  x  1000  =  17,384,150  foot-pounds. 
77.6 

Duty  per  million  B.T.U. 

=  1>349>Q4Q  x  1,000,000  =  16,893,863  foot-pounds. 
79,256 

Table  68  may  be  used  in  approximating  the  duty,  thus  : 

The  mechanical  efficiency  of  the  pump  in  the  preceding  problem  is 


PUMPS 


537 


X 


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538 


STEAM  POWER  PLANT  ENGINEERING 


At  the  intersection  of  vertical  column  "  85  "  and  horizontal  column 
100  of  Table  68,  we  find  16.82  millions. 

Tables  69  and  70  give  the  maximum  theoretical  height  to  which 
pumps  may  lift  water  by  suction  at  different  temperatures.  In  prac- 
tice these  figures  cannot  be  realized.  It  is  customary  to  have  the  water 
gravitate  to  the  pump  for  all  temperatures  over  120  degrees  F. 

TABLE   69. 

MAXIMUM  HEIGHT  TO  WHICH  PUMPS  CAN  RAISE  WATER  BY  SUCTION. 
(Temperature  of  Water  40  degrees  F.;  Barometer  29.92.) 


Vacuum  in 
Suction  Pipe, 
Inches  of 
Mercury. 

Theoretical 
Lift. 

Probable 
Actual 
Lift. 

Vacuum  in 
Suction  Pipe, 
Inches  of 
Mercury. 

Theoretical 
Lift. 

Probable 
Actual 
Lift. 

Feet. 

Feet. 

Feet. 

Feet. 

1 

1.1 

0.9 

16 

18.0 

14.4 

2 

2.2 

1.8 

17 

19.1 

15.3 

3 

3.3 

2.7 

18 

20.2 

16.1 

4 

4.5 

3.6 

19 

21.4 

17.1 

5 

5.6 

4.5 

20 

22.5 

18.0 

6 

6.7 

5.4 

21 

23.7 

18.9 

7 

7.9   . 

6.3 

22 

24.8 

19.8 

8 

9.0 

7.2 

23 

25.9 

20.7 

9 

10.1 

8.1 

24 

27.0 

21.6 

10 

11.3 

9.0 

25 

28.2 

22.7 

11 

12.4 

9.9 

26 

29.3 

23.9 

12 

13.5 

10.8 

*  27 

30.4 

24.3 

13 

14.6 

11.7 

28 

31.6 

25.2 

14 

15.8 

12.6 

29 

32.7 

26.1 

15 

16.9 

13.5 

f  29.68 

33.6 



*   Vacua  greater  than  27  inches  are  practically  unobtainable  in  pumping  practice  except  in 
connection  with  condensers. 

t  Maximum  theoretical  vacuum  obtainable  with  water  at  40  degrees  F.  and  barometer  of 
29.92  inches. 

TABLE   70. 

MAXIMUM   THEORETICAL   HEIGHT   TO  WHICH   A    PUMP   CAN   LIFT  WATER   BY 

SUCTION  AT  DIFFERENT  TEMPERATURES. 

(Barometer  29.92.) 


Temperature  of 
Feed  Water. 

Maximum 
Theoretical  Lift. 

Temperature  of  Feed 
Water. 

Maximum 
Theoretical  Lift. 

Degrees  F. 

Feet. 

Degrees  F. 

Feet. 

40 

33.6 

130 

29.2 

50 

33.5 

140 

27.8 

60 

33.4 

150 

25.4 

70 

33.1 

160 

23.5 

80 

32.8 

170 

20.3 

90 

32.4 

180 

16.7 

100 

31.9 

190 

12.8 

110 

31.3 

200 

7.6 

120 

30.3 

210 

1.3 

PUMPS  539 

275.   Size  of  Boiler-Feed  Pump. 

Let  D  =  diameter  of  water  cylinder,  inches. 

d  =  diameter  of  the  steam  cylinder,  inches. 

L  =  length  of  stroke,  inches. 

N  =  number  of  working  strokes  per  minute. 

H  =  head  in  feet  between  suction  and  boiler  water  level. 

R  =  resistance  in  pounds  per  square  inch  between  suction  level 
and  boiler  water  level  due  to  valves,  pipes,  and  fittings. 

p  =  boiler  pressure,  pounds  per  square  inch. 

S  =  ratio  of  the  water  actually  delivered  to  the  piston  displace- 
ment. 

W  =  weight  of  water  delivered,  pounds  per  hour. 

/  =  indicated  horse  power  of  the  pump  at  maximum  capacity. 

E  =  mechanical  efficiency  of  the  pump,  taken  as  the  ratio  of  the 
water  horse  power  at  the  discharge  opening  to  the  indicated 
horse  power  of  the  pump,  steam  end. 

Then 


°-433 


62.5  X  S  =  l.7D2LNS.  (132) 
(133) 

i:  (134) 


33,000  X  60  X  E 

In  average  practice  the  piston  or  plunger  displacement  is  made  about 
twice  the  capacity  found  by  calculation  from  the  amount  of  water 
required  for  the  engine,  to  allow  for  leakage,  steam  consumption  of  the 
auxiliaries,  blowing  off,  and  pump  slip. 

For  pumps  with  strokes  of  12  inches  or  over,  the  speed  of  the  plunger 
or  piston  is  usually  limited  to  100  feet  per  minute  as  a  maximum  to  insure 
smooth  running.  For  shorter  strokes  a  lower  limit  should  be  used.  The 
maximum  number  of  strokes  ranges  from  100  for  strokes  over  12  inches 
in  length,  to  200  for  strokes  under  5  inches.  Boiler-feed  pumps  should 
be  designed  to  give  the  desired  capacity  at  about  one-half  the  maximum 
number  of  strokes  or  less. 

Pump  slip  varies  from  5  to  40  per  cent,  depending  upon  the  condition 
of  the  piston  and  valves  and  the  number  of  strokes.  An  average  value 
for  piston  and  plunger  pumps  in  first-class  condition  is  8  per  cent  when 


540  STEAM  POWER  PLANT  ENGINEERING 

operating  at  rated  capacity,  but  it  is  wise  to  allow  a  much  larger  figure, 
say  20  per  cent,  for  leakage  caused  by  wear. 

The  area  of  the  steam  cylinder  is  made  from  2  to  2.5  times  that  of 
the  water  end  to  allow  for  the  various  friction  losses  and  the  drop  in 
pressure  between  the  pump  throttle  and  the  boiler.  The  total  head 
pumped  against  includes  the  suction  lift,  the  friction  of  valves  and 
fittings,  the  distance  between  the  suction  inlet  and  the  boiler  level,  and 
the  boiler  pressure.  The  excess  head  varies  in  practice  from  15  to  40 
per  cent  of  the  boiler  pressure;  an  average  figure  is  25  per  cent.  In 
allowing  for  the  drop  in  steam  pressure  between  boiler  and  pump  a  liberal 
figure  is  25  per  cent. 

The  application  of  formulas  (132)  to  (135),  including  the  practical 
considerations  stated  above,  is  best  illustrated  by  a  specific  example. 

Example:  Determine  the  size  of  direct-acting  single-cylinder  feed 
pump  necessary  to  supply  water  to  1000  horse  power  of  boilers.  Gauge 
pressure  100  pounds  per  square  inch;  feed- water  temperature  150 
degrees  F. 

One  horse  power  is  equivalent  to  the  evaporation  of  34.5  pounds  of 
water  from  and  at  212  degrees  F.;  but  the  pump  is  usually  designed  to 
supply  about  twice  the  capacity. 

Thus     W  =  62,400  (under  the  given  conditions). 

S  =  0.8  (by  assumption). 
LN  =  1200  (on  the  basis  of  100  feet  per  minute). 

Substitute  these  values  in  (133) : 

62,400  > '     call  it  6  i 


1200  X  0.8 

since  the  assumptions  have  been  very  liberal. 
Assume  (.433  H  +  R)  =  0.25  p  and  E  =  0.65. 
Substitute  these  values  in  (134) : 

d  =  a  <  :  100  +  25 


1.65  X  100 
=  8.35,  —  call  it  8.5  inches. 

Allowing  100  strokes  per  minute  the  length  of  the  stroke  must  be 
L  =  1200  ^  100  =  12  inches. 

The  dimensions  of  the  pump  are  8J  x  6  x  12. 

The  indicated  horse  power  at  maximum  load  may  be  obtained  by 
substituting  the  proper  values  in  (135),  thus: 

,  =       62,400  (100  +  25)  2.3 

33,000  X  60  X  0.65 
=  13.9  I.H.P. 


PUMPS 


541 


276.  Steam-Pump  Governors.  —  Fig.  276  shows  a  section  through  a 
Fisher  pump  governor,  illustrating  a  device  for  maintaining  a  practically 
constant  pressure  in  the  discharge  pipe  irrespective  of  the  quantity  of 
water  flowing.     It  embodies  a   pressure-reducing  valve  in  the  steam 

e  supply  pipe  of  the  pump,  actuated  by  the  slight 

variations  in  water  pressure.  When  the  demand 
for  water  increases,  the  pressure  in  the  discharge 
pipe  tends  to  decrease,  and  this  drop  in  pressure 
(transmitted  to  the  pump  governor  by  suitable 
piping)  causes  more  steam  to  be  admitted,  which 
increases  the  speed  of  the  pump.  The  governor 
is  connected  to  the  steam  inlet  of  the  pump  at 
B  and  the  steam  enters  at  A.  Double-balanced 
valve  C  regulates  the  supply  of  steam  to  the 
cylinder  by  the  amount  it  is  raised  from  the 
seat.  The  valve  is  held  open  by  spring  G,  the  com- 
pression of  which  may  be  regulated  by  hand  wheel 
K.  The  water  pressure  from  the  discharge  pipe 
acts  on  piston  F  and  tends  to  overcome  the  resist- 
ance of  the  spring.  The  difference  in  pressure 
between  the  water  and  the  spring  determines  the 
position  of  valve  C. 

Piston  rod  H  is  pinned  to  sleeve  I  and  valve 
stem  L  screwed  into  this  sleeve  by  means  of  hand 
wheel  K.  Hence  during  ordinary  operation  the 
piston,  piston  rod  sleeve,  valve  stem,  and  valve 
act  as  a  single  unit.  By  turning  the  hand  wheel 
K,  valve  stem  L  will  screw  into  sleeve  7  and  the 
tension  on  the  spring  will  be  increased.  Hand  wheel  J  serves  as  a 
lock  nut  and  prevents  K  from  turning  during  normal  operation. 

277.  Feed- Water  Regulators.  —  The  water  level  in  the  boiler  should 
be  kept  as  nearly  constant  as  possible,  and  this  necessitates  considerable 
attention  on  the  part  of  the  fireman,  especially  with  fluctuating  loads. 
There  are  a  number  of  devices  on  the  market  which  are  designed  to 
automatically  maintain  a  constant  level,  and  in  many  small  plants 
where  the  duties  of  the  fireman  are   numerous  such  devices  in  connec- 
tion with  high  and  low  water  alarms  are  of  considerable  assistance. 
Their  action,  however,  is  not  always  positive  on  account  of  wear  or 
sticking  of  parts,  and  engineers  as  a  rule  prefer  to  rely  upon  hand  regula- 
tion.    In  large  stations  regulators  are  seldom  used. 

Fig.  277  shows  a  section  through  a  Kitts  feed-water  regulator,  con- 
sisting of  two  parts,  the  chamber  F  and  the  regulating  valve  V,    The 


Fio.  276.    Fisher  Pump 
Governor. 


542 


STEAM  POWER  PLANT  ENGINEERING 


float  chamber  is  connected  to  the  boiler  or  water  column  at  0  and  E, 
and  the  regulating  valve  to  the  feed  main  at  R  and  to  the  boiler  feed 
pipe  at  W.  When  the  water  in  the  boiler  falls  below  the  mean  level,  the 
weight  B  overcomes  the  counterweight  G  and  closes  needle  valve  L  by 
means  of  compound  levers.  At  the  same  time  an  extension  on  valve  L 
lifts  spring  A  and  opens  exhaust  valve  D.  This  removes  the  steam 


FIG.  277.    Kitts  Feed-Water  Regulator. 


FIG.  278.    Rowe  Feed-Water 
Regulator. 


pressure  from  the  top  of  diaphragm  C,  in  the  regulating  valve,  through 
the  agency  of  pipe  K.  The  pressure  from  the  pump  raises  the  disk  T 
and  water  flows  into  the  boiler  until  the  water  rises  to  the  mean  level. 
When  weight  B  becomes  submerged  its  weight  is  overcome  by  counter- 
weight G,  valve  L  is  opened  and  exhaust  valve  D  is  closed.  This 
admits  steam  pressure  to  the  diaphragm  C  and  forces  disk  T  to  its  seat, 
cutting  off  the  supply  of  water  to  the  boiler. 

The  Rowe  feed-water  regulator,  Fig.  278,  depends  for  its  operation 
on  a  familiar  float-controlled  valve  mechanism.  The  vessel  A  is  con- 
nected to  the  boiler  above  and  below  the  water  line,  and  the  float  (7, 
following  the  water  level  up  and  down,  actuates  a  balanced  valve  in 
accordance  with  the  boiler-feed  requirements.  When  this  apparatus 
is  used  to  regulate  the  feed  of  a  single  boiler  the  opening  G  in  the  valve 
chamber  is  connected  to  the  steam  space  of  the  boiler  and  the  outlet  H 


PUMPS 


543 


is  carried  to  the  steam  inlet  of  the  feed-water  pump.  When  the  water 
level  is  normal  the  float  closes  the  valve  L  and  thereby  cuts  off  the 
supply  of  steam  to  the  pump  cylinders.  Communication  between 
chambers  A  and  R  is  prevented  by  means  of  a  diaphragm  M.  When 
the  water  level  falls  below  normal  the  float  pulls  the  valve  down,  open- 
ing the  way  for  steam  to  pass  from  the  inlet  G  to  the  outlet  H  and 
thence  to  the  pump.  When  the  regulator  is  used  to  control  a  battery 
of  boilers  the  pump  discharge  delivers  into  the  inlet  G  and  the  water 


FIG.  279.     A  Typical  Geared  Triplex  Pump. 

passes  through  H  to  the  boiler-feed  main.  Should  the  water  level  fall 
beyond  a  predetermined  limit  by  reason  of  any  accidental  discontinuance 
of  the  water  supply  which  the  apparatus  cannot  correct,  the  float  would 
open  the  valve  F  of  the  alarm  whistle  0  mounted  on  the  top  of  the 
main  vessel. 

278.  Power  Pumps.  —  Piston  pumps,  geared,  belted,  or  direct  con- 
nected to  electric  motors,  gas  engines,  and  water  motors,  are  used 
chiefly  where  steam  power  is  not  available.  Their  general  utility  is 
evidenced  by  the  rapidly  increasing  number  installed  in  situations 
formerly  occupied  by  the  direct-acting  steam  pump.  The  efficiency  of 


544 


STEAM  POWER  PLANT  ENGINEERING 


this  type  of  pump  depends  in  a  large  measure  upon  the  character  of 
the  driving  motor  and  the  efficiency  of  the  transmitting  mechanism. 
High-speed  power  pumps  direct  connected  to  electric  motors  give 


Knowles  High  Speed  Electric  Pump 

Direct  connected  to 

M  P  6-100  H.P.-280-220  V.Form  L 

Load  and  Efficiency 


Gauge  Pressure  at  Valve  (Lb.) 
Fia.  279a. 


efficiencies  from  line  to  water  horse  power  as  high  as  83  per  cent,  while 
the  low-speed  geared  type  seldom  exceed  70  per  cent.  The  curves  in 
Fig.  279a  give  the  performance  of  a  direct-connected  triplex  pump, 


PUMPS 


545 


and  those  in  Fig.  280  the  performance  of  a  triplex  pump  geared  to  an 
electric  motor.  Both  of  these  performances  are  exceptionally  good 
and  are  considerably  above  the  average. 

Power  Pumps:  Eng.  Mag.,  Jan.,  1905,  p.  616;  Col.  Guard,  Nov.  17,  1905;  Elec. 
Rev.,  Jan.  10,  1902;  Elec.,  N.Y.,  Oct.  12,  1904;  Elec.  World,  Oct.  14,  1905,  p.  667; 
Engr.,  Lond.,  Oct.  17,  1902,  p.  377;  Engineering,  Sept.  1,  1905,  p.  275;  Engr.  U.S., 
Jan.  1,  1904,  p.  47. 


400  600  800  1000 

Discharge,  Gallons  Per  Minute 

Head  Constant.     Speed  Variable. 


1200 


1400 


tor  K.P.M.  450 
Gearing  10  to  1 
Capacity  1250  Gal.  per  Min. 


50  75  100  125 

Total  Head,  Lb.  Per  Sq.  In.  Gauge 
Speed  Constant.     Head  Variable. 

FIG.  280.    Performance  of  a  65-Horse-Power,  Motor-Driven  Triplex  Pump.    Geared  Type. 

279.  Injectors.  —  As  a  boiler  feeder  the  injector  is  an  efficient 
and  convenient  device,  cheap  and  compact,  with  no  moving  parts, 
delivers  hot  water  to  the  boiler  without  preheating,  and  has  no  exhaust 


546 


STEAM  POWER  PLANT  ENGINEERING 


steam  to  be  disposed  of.  Its  adoption  in  locomotives  is  practically 
universal,  but  in  stationary  practice  it  is  limited  to  small  boilers  or 
single  boilers  or  as  a  reserve  feeder  in  connection  with  pumps.  The 
objections  to  an  injector  are  its  inability  to  handle  hot  water,  the 
difficulty  of  maintaining  a  continuous  flow  under  extreme  variation  of 
load,  and  the  uncertainty  of  operation  under  certain  conditions.  Fig. 
281  illustrates  the  simplest  form  of  single-tube  injector.  Boiler  steam 


FIG.  281.    Elementary  Steam  Injector. 

is  admitted  at  A  and,  flowing  through  nozzle  and  combining  tube  to 
the  atmosphere  through  G,  partially  exhausts  the  air  from  pipe  B, 
thereby  causing  the  water  to  rise  until  it  comes  in  contact  with  the  steam. 


Steam 


Overflow 

FIG.  282.     Hancock  Double-Tube 
Injector. 


FIG.  283.     Penberthy  Automatic 
Injector. 


The  steam  emerging  from  nozzle  C  at  high  velocity  condenses  on  meeting 
the  water  and  imparts  considerable  momentum  to  it.     The  energy  in 


PUMPS  547 

the  rapidly  moving  mass  is  sufficient  to  carry  it  across  opening  0,  lift 
check  H  from  its  seat  and  force  it  into  the  boiler.  The  steam  then 
ceases  to  escape  at  G. 

280.  Positive  Injectors.  —  Fig.  282  shows  a  section  through  a  Han- 
cock injector,  illustrating  the  principles  of  the  double-tube  positive 
type.      Its   operation   is    as  follows:    Overflow  valves   D  and  F  are 
opened  and  steam  is  admitted,  which  at  first  passes  freely  through  the 
overflow  to  the  atmosphere  and  in  so  doing  exhausts  the  air  from  the 
suction  pipe.     This  causes  the  feed  water  to  rise  until  it  meets  the  jet 
of  steam  and  the  two  are  forced  through  the  overflow.     As  soon  as 
water  appears  at  the  overflow,  valve  D  is  closed,  valve  C  partially 
opened,  and  valve  F  closed.     This  admits  steam  through  the  forcing 
jet  W  and,  the  overflow  valves  being  closed,  the  water  is  fed  into  the 
boiler.     In  case  the  action  is  interrupted  for  any  reason  it  is  necessary 
to  restart  it  by  hand. 

The  chief  advantage  of  the  double-tube  positive  type  lies  in  its  ability 
to  lift  water  to  a  greater  height  and  to  handle  hotter  water  than  the 
single-tube.  Its  range  in  pressure  is  also  greater,  that  is,  it  will  start 
with  a  lower  steam  pressure  and  discharge  against  a  higher  back  pressure. 
Double-tube  injectors  are  used  almost  exclusively  in  locomotive  work. 

281.  Automatic  Injectors.  —  Fig.  283  shows  a  section  through  the 
Penberthy  injector.      Its  operation  is  as  follows:   Steam  enters  at  the 
top  connection  and  blows  through  suction  tube  c  into  the  combining 
tube  d  and  into  chamber  g,  from  which  it  passes  through  overflow  valve  n 
to  the  overflow  m.     When  water  is  drawn  in  from  the  suction  intake 
and  begins  to    discharge  at  the  overflow,  the  resulting  condensation 
of  the  steam  creates  a  partial  vacuum  above  the  movable  ring  h  and 
the  latter  is  forced  against  the  end  of  tube  c,  cutting  off  the  direct  flow 
of  water  to  the  overflow.     The  water  then  passes  into  the  boiler.     Spill 
holes  i,  i,  i  are  for  the  purpose  of  relieving  the  excess  of  water  until 
communication  with  the  boiler  has  been  established.     The  action  of 
opening  and  closing  the  overflow  is  entirely  automatic.     Where  the 
conditions  are  not  too  extreme  the  automatic  injector  is  to  be  preferred 
for  stationary  work  because  of  its  restarting  features.     It  is  also  used  on 
traction,  logging,  and  road  engines,  where  its  certainty  of  action  and 
special  adaptability  render  it  invaluable  for  the  rough  work  to  which 
such  machines  are  subjected. 

Injectors,  Theory  of:  Trans.  A.S.M.E.,  10-339;  Sibley  Jour.,  Dec.,  1897,  p.  101; 
Power,  May,  1901,  p.  23;  Thermodynamics  of  the  Steam  Engine,  Peabody,  Chap. 
IX;  Theory  of  the  Steam  Injector,  Kneass. 

Injectors,  General  Description:  Engr.  U.S.,  Oct.  1,  1907,  Nov.  15,  1907,  July  15, 
1904,  p.  501,  Feb.  2,  1903,  p.  151;  Power,  Aug.,  1906,  p.  478;  Engr.,  Lond.,  March 
10,  1905,  p.  244;  Engineering,  Aug.  30,  1895,  p.  281. 


548  STEAM  POWER  PLANT  ENGINEERING 

282.  Performance  of  Injectors.  —  The  performance  of  an  injector 
may  be  very  closely  determined  from  the  equation 

_  xr  +  q  —  t  +  32    (Kneass,  "  Theory  of  the 

t-t.  Injector/' p.  83). 

in  which 

w   =  pounds  of  water  delivered  per  pound  of  steam  supplied. 

x   =  quality  of  the  steam  supplied. 

r   =  heat  of  vaporization. 

q    =  heat  of  the  liquid. 

t   =  temperature  of  the  discharge  water. 

IQ  =  temperature  of  the  suction  water. 

Figs.  284a,  284b,  and  284c  give  the  performance  of  a  Desmond  auto- 
matic injector  as  tested  at  the  Armour  Institute  of  Technology.  The 
results  check  very  closely  with  those  calculated  from  above  equation. 
Referring  to  Fig.  284a  it  will  be  seen  that  the  weight  of  water  delivered 
per  pound  of  steam  decreases  as  the  initial  pressure  is  increased,  all 
other  factors  remaining  the  same.  From  Fig.  284b  it  will  be  noted 
that  the  weight  of  water  delivered  per  pound  of  steam  decreases  as  the 
temperature  of  suction  supply  is  increased  up  to  a  point  where  the 
injector  "  breaks  "  or  becomes  inoperative.  This  critical  temperature 
varies  with  the  different  types  of  injectors,  being  highest  for  the 
double-tube  type,  but  seldom  exceeds  160  degrees  F.  Fig.  284c  shows 
that  the  weight  of  water  delivered  per  pound  of  steam  is  practically 
constant  for  all  discharge  pressures  within  the  limits  of  the  apparatus. 

Table  71  gives  the  range  of  working  steam  pressures  for  standard 
"  Metropolitan "  injectors  with  varying  suction  heads  and  temper- 
atures, and,  though  strictly  applicable  to  this  particular  type  only, 
is  characteristic  of  all  makes. 

In  selecting  an  injector  the  following  information  is  desirable  for 
best  results: 

1.  The  lowest  and  highest  steam  pressure  carried. 

2.  The  temperature  of  the  water  supply. 

3.  The  source  of  water  supply,  whether  the  injector  is  used  as  a 
lifter  or  non-lifter. 

4.  The    general    service,    such    as    character   of    the   water    used, 
whether  the  injector  is  subject  to  severe  jars,  etc. 

Injectors,  Tests  of:  Eng.  News,  March  17,  1898,  July  16,  1896,  p.  39;  Locomotive 
Engineering,  May,  1900,  p.  204;  Power,  Oct.,  1904,  p.  602;  Railroad  Gazette,  Dec. 
11,  1896;  Thermodynamics  of  the  Steam  Engine,  Peabody,  Chapter  IX;  Theory  of 
the  Injector,  Kneass. 


PUMPS 


549 


i  a 


c 

) 

\ 

Constant  Discharge  Pressure 
20  Lb.  Per  Sq.In. 
Constant  Suction  Temp. 
55  Deg.Fah. 

\ 

\ 

^s 

\ 

xc 

\ 

^ 

\, 

^ 

) 

^ 

^ 

65  70  75  80  85  90 

Initial  Gauge  Pressure.Lb.Per  Sq.In. 


95 


FIG.  284a.    Performance  of  an  Automatic  Injector  with  Varying  Initial  Pressure. 


I 

la  I8 


|4u 

5      *5 


Constant  Initial  Pressure 

70  Lb.  Per  Sq.  In. 

Constant  Discharge  Pressure 

70  Lb.  Per  Sq.  In. 


75  85  95 

Temperature  of  Suction,  Deg.  Fan. 


105 


115 


FIG.  284b.     Performance  of  an  Automatic  Injector  with  Varying  Suction  Temperature 


Lb.  'Water  Delivered 
Per  Lb.  of  Steam 

a  s  s  s  & 

Constant  Initial  Pressure,70  Lb.  Per  Sq.  In. 
Constant  Suction  Temperature,  56  Deg.  Fan. 

( 

( 

j 

; 

20                 30                 AO                50                  60                  70                 QQ 

Discharge.PressuEe,.Lb.  PetSq.  In.  Gauge 


FIG.  284c.    Performance  of  an  Automatic  Injector  with  Varying  Discharge  Pressure. 


550 


STEAM  POWER  PLANT  ENGINEERING 


TABLE  71. 

RANGE  IN  WORKING  PRESSURES. 

Standard  "  Metropolitan  "  Steam  Injectors. 


Automatic. 

Suction 
Temperature, 

I 

Suction  Head,  P 

eet. 

Degrees  F. 

2 

8 

14 

20 

Under 
Pressure. 

Under  60 

25  to  150 

30  to  130 

42  to  110 

55  to    85 

20  to  160 

100 

26  to  120 

33  to  100 

55  to    80 

25  to  125 

120 

26  to    85 

140 

Suction 
Temperature, 
Degrees  F. 

Double  Tube. 

Suction  Head,  Feet. 

2 

8 

14 

20 

Under 
Pressure. 

Under  60 

14  to  250 

23  to  220 

27  to  175 

42  to  135 

14  to  250 

100 

15  to  210 

26  to  160 

37  to  120 

46  to    70 

15  to  210 

120 

20  to  185 

30  to  120 

42  to    75 

20  to  185 

140 

20  to  120 

35  to    70 

20  to  120 

283.  Injector  vs.  Steam  Pump  as  a  Boiler  Feeder. —  From  a  purely 
thermodynamic  standpoint  the  efficiency  of  an  injector  is  nearly 
perfect,  since  the  heat  drawn  from  the  boiler  is  returned  to  the  boiler 
again,  less  a  slight  radiation  loss.  As  a  pump,  however,  the  injector  is 
very  inefficient  and  requires  more  fuel  for  its  operation  than  very 
wasteful  feed  pumps.  This  is  best  illustrated  by  an  example:  An 
injector  of  modern  construction  will  deliver  say  15  pounds  of  water  to 
the  boiler  per  pound  of  steam  supplied,  with  delivery  temperature  of 
150  degrees  F.  This  corresponds  to  a  heat  consumption  of  71  B.T.U. 
per  pound  of  water  delivered,  thus : 

With  initial  pressure  of  115  pounds  absolute, 

X  =  1185. 


PUMPS  551 

Heat  in  the  water  delivered  to  the  boiler, 

150  -  32  =  118  B.T.U.  above  32  degrees  F. 
Heat  of  1  pound  of  steam  above  a  feed  temperature  of  150  degrees  F. 

1185  -118  =  1067  B.T.U. 
Heat  required  to  deliver  1  pound  of  water  to  the  boiler, 


=  71  B.T.U. 
lo 

A  simple  direct-acting  duplex  pump  consumes  say  200  pounds  steam 
per  I.H.P.  hour.  Assume  the  extreme  case  where  the  exhaust  steam 
will  not  be  used  for  heating  the  feed  water  and  the  latter  is  fed  into 
the  boiler  at  60  degrees  F. 

The  heat  supplied  to  the  pump  per  I.H.P.  hour, 

200  {1185  -  (60  -  32)}  =  231,400  B.T.U. 

Assuming  the  low  mechanical  efficiency  of  50  per  cent,  the  heat 
required  to  develop  one  horse  power,  at  the  water  end  will  be 

231,400  -5-  0.50  =  462,800  B.T.U.  per  hour. 

Since  the  steam  pressure  is  100  pounds  gauge,  the  equivalent  head 
of  water  at  60  degrees  F.  is 

2.3  X  100  =  230  feet. 

Assume  the  friction  in  the  feed  pipe,  the  resistance  of  valves,  etc.,  to 
be  30  per  cent  of  the  boiler  pressure;  the  total  head  pumped  against 
will  be 

230  +  69  =  299,  say  300  feet. 
1  horse-power  hour  =  1,980,000  foot-pounds  per  hour. 

1,980,000      „„ 

3Qp  -  =  6600  pounds, 

that  is,  1  horse  power  at  the  pump  will  deliver  6600  pounds  of  water  per 
hour  to  the  boiler  against  a  head  of  300  feet. 

The  heat  consumption  per  pound  of  water  delivered, 


If  the  feed  water  is  heated  to  say  210  degrees  F.  by  the  exhaust  steam 
from  the  pump,  the  heat  consumption  will  be  55.5  B.T.U.  as  against  70.1 
without  the  heater. 

Thus  even  in  this  extreme    case  of  poor  steam-pump  performance 
the  heat  consumption  lies  in  favor  of  the  pump.     With  the  better 


552 


STEAM  POWER  PLANT  ENGINEERING 


grades  of  pumps  this  disparity  is  considerably  greater,  and  decidedly 
so  if  the  exhaust  steam  is  used  to  preheat  the  feed  water.  For  inter- 
mittent operation  the  condensation  losses  in  the  pump  may  more 
than  offset  this  gain.  Other  conditions,  however,  such  as  compact- 
ness, low  first  cost,  and  ease  of  operation  are  oftentimes  considerations, 
and  the  heat  consumption  is  of  minor  importance. 
284:.  Air  Pumps.  —  Condenser  air  pumps  may  be  divided  into  two 


1.  Wet-air  pumps  and 

2.  Dry-air  pumps. 

The  former  handle  both  air  and  water  and  the  latter  air  alone. 
Ordinary  jet-condenser  wet-air  pumps  handle  simultaneously  the 
circulating  water,  condensed  steam,  and  entrained  air,  and  are,  in 
fact,  a  combination  of  circulating  and  vacuum  pump.  Surface- 
condenser  wet-air  pumps  are  the  same  in  principle  and  design,  but 
are  smaller  in  size  for  a  given  main  engine  output,  as  they  handle  the 
condensed  steam  and  air  only. 

Wet-air  pumps  may  be  driven  by  the  main  engine  or  independently 
and  may  be  direct  acting,  Fig.  203,  or  fly  wheel  driven,  Fig.  228. 
The  fly-wheel  type  may  be  steam,  electric,  or  belt  driven.  Dry-air 

pumps  are  virtually  air  compres- 
sors, as  their  function  is  to  com- 
press air  from  the  pressure  existing 
in  the  condenser  to  that  of  the 
atmosphere.  They  are  generally 
of  the  fly-wheel  type.  Where  a 
high  degree  of  vacuum  is  neces- 
sary the  air  cylinders  are  com- 
pounded, as  efficient  compression 
of  say  J  pound  to  15  pounds  or  30 
to  1  is  too  great  for  a  single  stage. 
285.  Dean  Air  Pump.  —  Fig.  285 
shows  a  section  of  the  air  cylinder 
of  a  Dean  twin-cylinder  wet-air 
pump  as  applied  to  a  jet  condenser. 
There  are  three  sets  of  valves,  the 
suction  or  foot  valves  A,  A,  the  lift- 
FIG.  285.  Dean  Air  Pump.  inS  or  bucket  valves  B,  B,  and  the 

head  or  discharge  valves  (7,  C.    On 

the  upward  stroke  of  the  piston  or  bucket  a  partial  vacuum  is  formed  in 
the  chamber  between  the  bucket  and  the  lower  head,  causing  the  water 


PUMPS  553 

and  air  in  the  bottom  of  the  barrel  to  lift  the  foot  valves  A,  A  from 
their  seats  and  flow  into  the  cylinder.  On  the  downward  stroke  the  foot 
valves  A,  A  close  and  water  and  air  are  entrapped  in  chamber  R  between 
the  lower  head  and  the  bucket.  As  the  bucket  descends,  the  pressure 
of  air  in  the  cylinder  lifts  the  bucket  valves  B,  B  from  their  seats  and 
permits  the  air  and  water  to  escape  to  the  upper  portion  S  of  the 
cylinder  between  the  head  plate  and  the  bucket.  On  the  next  upward 
stroke  the  water  and  air  are  forced  through  the  discharge  valves  C,  C 
into  the  hot  well.  This  discharge  of  water  and  air  from  the  top  com- 
partment is  simultaneous  with  influx  of  water  and  air  in  the  lower 
chamber. 

See  paragraph  209  for  other  types  of  wet-air  pumps  in  connection 
with  jet  condensers. 

286.  Size  of  Wet-Air  Pumps;  Jet  Condensers.  —  In  proportioning 
such  pumps  the  quantity  of  cooling  water  and  condensed  steam  to  be 
taken  care  of  is  readily  determined,  but  the  percentage  of  air  mingled 
with  it  must  be  estimated.  Surface  water  under  atmospheric  pressure 
ordinarily  contains  from  2  to  5  per  cent  of  air  by  volume.  To  pro- 
vide for  possible  leakage  a  very  liberal  factor  is  usually  allowed,  an 
average  figure  being  about  10  per  cent. 

Let  Q  =  total  volume  of  air  and  water  in  cubic  feet  per  hour  to 

be  handled  by  the  pump. 

V  =  volume  of  cooling  water  in  cubic  feet  per  hour. 
v  =  volume  of  condensed  steam  in  cubic  feet  per  hour. 
va  =  volume  of  air  at  pressure  Pa  and  temperature  Ta. 
V  +  v  =  total  volume  of  water  and  condensed  steam  at  atmospheric 

temperature. 

Ta  =  temperature  of  the  air  entering  the  condenser,  degrees  F. 
T2  =  temperature  of  the  discharge  water,  degrees  F. 
TQ  =  initial  temperature  of  the  cooling  water,  degrees  F. 
Pa  =  atmospheric  pressure,  pounds  per  square  inch. 
Pc  =  total  pressure  in  the  condenser,  pounds  per  square  inch. 
Pv  =  pressure  of  aqueous  vapor  at  temperature  T2. 

Then  (V  +  v)  =  volume  of  water  to  be  pumped  from  the  condenser 
per  hour  and  va  =  volume  of  air  at  atmospheric  pressure  and  tempera- 
ture Ta  entering  the  condenser;  but  on  entering  the  condenser  the  air  is 
increased  in  volume,  due  to  the  reduction  in  pressure  and  the  increase 
in  temperature,  and  the  total  volume  to  be  exhausted  per  hour  by  the 
pump  is 


554 


STEAM  POWER  PLANT  ENGINEERING 


Under  average  conditions  of  reciprocating  engine  practice  v  =  ?V  V, 
Pa  =  15  pounds  per  square  inch,  Pc  =  2  pounds  per  square  inch, 
Pv  =  1.27,  T2  =  110  degrees  F.,  T0  =  60  degrees  F.  Substituting 
these  values  in  above  equations,  assuming  va  =  10%  of  V ,  we  get 
Q  =  3.4  V. 

Average  practice  gives  3  F  as  the  pump  displacement  per  hour  for 
a  single-acting  pump  and  3.5  V  for  a  double-acting  pump,  the  cylinders 
being  ordinarily  proportioned  on  a  basis  of  50  feet  per  minute  piston 
velocity  at  rated  capacity. 

Table  72  gives  the  approximate  sizes  of  air  pumps  for  condensers  as 
manufactured  by  prominent  makers. 

TABLE  72. 

APPROXIMATE   SIZES  OF  AIR  PUMPS  FOR  CONDENSERS. 


Jet  Condenser. 

Surface  O 

andensers. 

Steam  Con- 
densed per 
Hour. 

Duplex 
Pump. 

Horizontal 
Double  Acting 
Pump. 

Vertical 

2-Cylinder 
Single 
Acting. 

Horizontal. 

Vertical 
2-Cy  Under. 

500  to     1,000 

4f  X  5 

6X7 

5X4 

3*  X     4 

1,000  to     1,500 

5|  X  6 

8X7 

6X    4 

4X4 

1,500  to     2,000 

6i  X  6 

8X  12 

7X    5 

4X6 

2,000  to     2,500 

7iX  6 

9X9 

9X    6 

5X7 

2,500  to     3,000 

7    X  10 

9X  10 

10  X    8 

5X8 

3,500  to     4,000 

8    X  10 

11  X  12 

11  X    9 

6X8 

4,000  to     4,500 

8J  X  10 

12  X  14 

12  X    8 

7X9 

4,500  to     5,000 

9    X  10 

14  X  14 

12  X  10 

7    X  10 

5,000  to     6,000 

10    X  10 

14  X  16 

14  X  10 

7    X  12 

6,000  to     7,000 
7,000  to     8,000 
8,000  to     9,000 
9,000  to  10,000 
10,000  to  15,000 
15,000  to   20,000 
20,000  to  25,000 

10*  X  10 
11  X  10 
12  X  10 
12  X  15 
15  X  15 
17  X  15 
19  X  15 

15  X  16 
15  X  18 
16  X  18 
18  X  18 
20  X  24 
24  X  24 
26  X  24 

15  X  10 
15  X  12 
16  X  10 
17  X  12 
20  X  12 
22  X  15 
24  X  18 

8    X  10 
8    X  12 
9    X  12 
10    X  12 
12    X  14 
14    X  16 
16    X  24 

8X4 
8X    6 
9X    6 
10  X    8 
11  X    8 
12  X    8 
14  X  10 

Wet-air  pumps  are  usually  independently  driven,  making  it  possible 
to  vary  the  speed  of  the  pump  irrespective  of  the  engine  speed  and  to 
create  a  vacuum  before  starting  the  engine.  Occasionally,  however, 
when  the  load  is  constant,  as  in  pumping-engine  practice,  the  pump 
may  be  driven  by  the  main  engine. 

Air  Pumps:  Power,  Nov.,  1904,  p.  652;  Engineering,  Nov.  17,  1905;  Engr.  U.S., 
Jan.  1,  1906,  Elec.  Engr.,  Lond.,  Nov.  14,  1902;  American  Mach.,  Jan.  31,  1901, 
p.  113;  National  Engr.,  June,  1906. 

Air  Pumps.  —  Reference  Books:  Hausbrand,  Evaporating  and  Condensing 
Apparatus,  p.  383;  Whitham,  The  Steam  Engine,  p.  301;  Thurston,  Manual  of  the 
Steam  Engine,  Vol.  II,  p.  145;  Seaton,  Manual  of  the  Marine  Engine,  p.  301. 


PUMPS 


555 


287.  Edwards  Air  Pump.  —  Fig.  286  shows  a  section  through  the 
air  cylinder  of  an  Edwards  air  pump.  This  device  belongs  to  the  sur- 
face-condenser "  wet-air  pump  "  class,  as  both  the  water  of  condensa- 
tion and  the  entrained  air  are  exhausted  simultaneously  by  the  same 
piston.  Unlike  the  standard 
type  of  wet-air  pumps,  foot 
valves  and  bucket  valves  are 
entirely  dispensed  with.  The 
condensed  steam  flows  contin- 
uously by  gravity  from  the 
condenser  into  the  base  of  the 
pump  through  passage  A  and 
annular  space  B.  As  the  piston 
C  descends  it  forces  the  water 
from  the  lower  part  of  the 
casing  F  into  the  cylinder 
proper  through  the  ports  P,  P. 
On  the  upward  stroke  the 
ports  in  the  piston  are  closed 
and  the  air  and  water  dis- 
charged through  head  valves 
D  and  exhaust  port  E  to 
the  hot  well.  The  seats  of 


FIG.  286.    Edwards  Air  Pump. 


valves  D  are  constructed  with  a  rib  between  each  valve  and  a  lip 
around  the  outer  edge,  so  that  each  valve  is  water-sealed  independently, 
of  the  others.  In  earlier  air  pumps  of  this  general  type  the  clearance 
between  the  bucket  and  head  valve  seat  is  necessarily  large,  due  to 
the  space  occupied  by  the  bucket  valves  and  the  ribs  on  the  under  side 
of  the  valve  seating.  This  clearance  space  reduces  the  capacity  of  the 
pump,  since  the  air  above  the  bucket  must  be  compressed  above 
atmospheric  pressure  before  it  can  be  discharged,  and  on  the  return 
stroke  will  expand  and  occupy  a  space  which  should  be  available  for  a 
fresh  supply  of  air  from  the  condenser.  In  the  Edwards  air  pump  the 
clearance  space  is  reduced  to  a  minimum,  since  there  are  no  bucket 
valves  to  limit  it.  The  absence  of  suction  or  foot  valves  still  further 
increases  the  capacity  of  the  pump  for  similar  reasons.  These  pumps 
are  arranged  either  single,  double,  or  triplex;  steam,  electric,  or  belt 
driven;  slow  or  high  speed.  They  are  ordinarily  used  in  connection 
with  surface  condensers. 

Edwards  Air  Pump:  Engr.,  July  1,  1903,  p.  536;  Engineering,  62-221,  63-60, 
64-767,  80-328;  Eng.  News,  June  12,  1902,  p.  478. 

Centrifugal  Wet  Vacuum  Pump:  Power  &  Engr.,  Jan,  4/1910. 


556 


STEAM  POWER  PLANT  ENGINEERING 


288.   Mullan  Valveless  Air  Pump.  —  Fig.  287  gives  several  views  of 
the  "  Mullan  valveless   air  pump "  as   used  in  connection  with  the 


Elevation  on 
Line  A-B 

Half  Section  on 
C.L.  of  Air  Cylinder 


FIG.  287.    Mullan  Valveless  Dry-Air  Pump. 


C.  H.  Wheeler  Company's  "  high-vacuum  "    condensing  outfit.     The 
pump  is  double  acting  and  devoid  of  suction  valves.     The  cylinder 


EXHAUST 


CONDENSER 


MAIN     PUMP  DRY    AIR   PUMP 

FIG.  288.     Hewes  and  Phillips  Air  Pumps. 


has  a  central  port  which  is  uncovered  by  the  piston  at  each  end  of 
the  stroke   and  covered  at  all  other  positions.     Discharge  valves  of 


PUMPS 


557 


the  familiar  spring-seated  poppet  type  are  located  in  both  heads  of  the 
cylinder.  As  the  piston  moves  from  one  end  of  its  stroke  to  the 
other  it  forms  a  vacuum  behind  it  and  forces  out  the  gases  and  water 
ahead  of  it;  when  it  reaches  the  end  of  the  stroke  the  central  inlet 
port  is  uncovered  and  the  vacuum  behind  the  piston  draws  in  the 
condensation  and  gases  from  the  condenser.  This  operation  is  repeated 
on  the  return  stroke. 

The  makers  claim  that  the  pump  will  operate,  under  shop  test 
conditions,  within  one-half  inch  of  the  barometer,  enabling  them  to 
guarantee  a  vacuum  within  two  inches  of  absolute  under  full-load 
conditions  of  steam  turbine  operation. 

289.  Alberger  Rotative  Dry-Air  Pump.  —  Fig.  289  shows  a  section 
through  the  air  cylinder  of  an  Alberger  rotative  dry-air  pump,  illus- 


FIG.  289.    Alberger  Rotative  Dry-Air  Pump. 

trating  a  type  of  pump  in  which  the  admission  valve  is  mechanically 
operated.  This  pump  is  designed  to  operate  with  dry  air  only,  all 
condensation  being  removed  before  the  air  enters  the  cylinder.  This 
permits  of  the  use  of  a  small  clearance  space  and  makes  it  possible  to 
run  at  a  higher  speed  of  rotation  than  can  be  secured  with  a  type  of 
pump  in  which  water  is  used  to  seal  the  valves.  Referring  to  Fig.  289, 
air  is  being  taken  into  the  right-hand  end  of  the  cylinder  through  inlet  A 
and  forced  from  the  left-hand  end  through  exhaust  opening  B.  Rotary 
valve  0  mechanically  opens  to  admission  and  mechanically  closes  the 
discharge.  The  discharge  opening  depends  on  the  spring-regulated  valve 
C  at  the  top  of  the  cylinder.  Heads  are  water  jacketed.  Ports  and 
passages  are  made  large  to  reduce  the  friction  of  the  air  entering  the 
pump,  and  to  obviate  the  bad  effects  of  clearance,  an  equalizing  passage 


558  STEAM  POWER  PLANT  ENGINEERING 

is  provided  in  valve  0.  The  action  of  the  passage  is  shown  in  the 
section  to  the  right.  When  the  piston  reaches  the  end  of  the  stroke 
the  clearance  space  is  filled  with  air  at  atmospheric  pressure.  If  this 
pressure  were  not  relieved  the  piston  would  travel  a  considerable  dis- 
tance before  drawing  in  air  from  the  condenser.  By  means  of  the 
equalizing  passage  the  clearance  space  is  connected  to  the  opposite  end 
of  the  cylinder  and  the  vacuum  there  reduces  the  pressure  in  the  clear- 
ance space. 

390.  Size  of  Wet-Air  Pump  for  Surface  Condenser.  —  Since  the  wet- 
air  pump  handles  both  the  air  and  condensed  steam,  its  theoretical 
capacity,  neglecting  clearance,  may  be  determined  by  eliminating  V 
from  equation  (138),  which  then  becomes 

- 

(139) 


For  the  average  reciprocating  engine,  Pc  =  2,  T2  —  Ta  =  110, 
Tn  =  60,  Pv  =  1.27.  Assuming  Pa  =  15  pounds  per  square  inch,  and 
substituting  these  values  in  (139),  Q  =  v  4-  25  va. 

The  volume  of  air  entering  the  condenser  varies  so  much  with  the 
character  of  the  power  plant  equipment  and  the  conditions  of  operation 
that  any  assumed  "  average  "  value  of  va  may  lead  to  serious  error. 
(See  Power  &  Engr.,  Feb.  2,  1909,  p.  234.) 

A  study  of  some  two  hundred  condenser  installations  gives 

Q  =  10  v  for  the  average  reciprocating  engine  and 
Q  =  20  v  for  average  steam  turbine  practice. 

Table  72  gives  the  cylinder  dimensions  of  wet-air  pumps  as  advo- 
cated by  prominent  condenser  builders. 

291.  Size  of  Dry-Air  Pumps.  —  "  Dry-air  "  pumps  are  used  in  con- 
nection with  barometric  and  surface  condensers  where  a  high  degree 
of  vacuum  is  essential,  as  in  steam  turbine  practice.  Such  pumps  are 
intended  to  exhaust  the  saturated  non-condensable  vapors  only. 

The  capacity  of  the  dry-air  pump  is  based  upon  experience  rather 
than  theory.  An  investigation  of  some  fifty  installations  gives 

Q  =  20  v  to  30  v  for  vacua  under  27  inches. 

Q  =  50  v  for  vacua  of  28  inches  and  over,  both  referred  to  a  30-inch 
barometer. 

Professor  Weighton  states  that  "  with  suitable  condenser  arrange- 
ments and  a  reasonably  air-tight  system  there  is  nothing  gained  in 
efficiency  by  the  use  of  air  pumps  exceeding  in  capacity  0.7  of  a  cubic 
foot  per  pound  of  steam  condensed  up  to  a  vacuum  of  29  inches." 
(Engineering  Record,  May  19,  1906,  p.  61.) 


PUMPS  559 

The  work  done  by  the  average  "  high- vacuum  "  dry-air  pump  is  a 
maximum  for  vacua  between  18  and  20  inches. 

This  may  be  proved  from  Fig.  290,  which  represents  a  theoretical 
indicator  card  from  the  air-pump  cylinder. 

Let  p2  =  pressure  in  the  condenser,  pounds  per  square  inch  absolute. 
pl  =  atmospheric  pressure. 

v2  =  piston  displacement,  including  clearance,  cubic  feet. 
v^  =  volume  of  air  in  the  cylinder  when  the  valve  opens   to 

atmosphere,  cubic  feet. 
vc  =  clearance  volume,  cubic  feet. 

The  work  done  is  proportional  to  the  area  A  BCD. 
Area  ABCD  =  work  done  =  area  EBIO  +  BAGI  -FAGO  -  ECDF. 
Neglecting  the  exponential  factor  n  for  the  sake  of  simplicity,  thus 
making  pv  =  p1vl  =  p2v2  =  constant, 

W  =  work  done  =  plvl  +  plvl  I     — —  p2v2  •—  pjVe  +  p2ve.        (140) 

Jv2      V 

Substitute  p^  for  its  equivalent  p2v9  and  PiviVc  for  its  equiva- 
lent p2ve  and  integrate. 

W  =  p1vl  +  plvl  log,^2  —  plvl  log,  vl  —  plvl  +       l  c>  (141) 

making  the  first  derivative  zero. 

^  =  0  =  log,  t>2  -  1  -  log,  Vl  +  ^  (142) 

0  =  log  y2-!  +%,  (143) 

v\  vi 

i.e.,  W  is  a  maximum  when 


Y)  9) 

or  log,  —  =1  --  -,  since  plvl  =  p2v2.  (145) 

P2  V2 

For  average  high-vacuum  practice  vc  =  3  per  cent  of  the  piston  dis- 
placement. Assume  v2=  1,  vc  =  0.03,  and  p^  =  14.7  pounds  per  square 
inch,  and  substitute  these  values  in  (145),  thus: 

14.7  0.03 


Whence  p2  =  5.5  pounds  per  square  inch  absolute,  which  corresponds 
to  a  vacuum  of  18.6  inches  of  mercury. 


560 


STEAM  POWER  PLANT  ENGINEERING 


Thus  we  see  that  the  maximum  load  on  the  pump  occurs  when  the 
vacuum  is  between  18  and  20  inches.  If  the  vacuum  is  less  than  this, 
the  load  falls  off  because  of  the  decreased  difference  in  pressure.  If  the 


FIG.  290. 

vacuum  is  greater,  the  load  falls  from  the  decrease  in  weight  of  air 
handled. 

291a.  Hot-well  Pumps.  —  In  high-vacuum  surface  condensers  the 
condensed  steam  is  often  handled  by  a  small  independently  driven  pump 
called  the  hot-well  pump.  See  paragraph  228. 

Types  of  Air  Pumps:  Power  &  Engr.,  June  1,  1909,  p.  963. 

292.  Centrifugal  Pumps.  —  Centrifugal  pumps  consist  of  two  essen- 
tial elements,  (1)  a  rotating  impeller  which  draws  in  the  water  at  its 
center,  and  (2)  a  stationary  casting  which  guides  the  water  thrown 


Types  of  Impellers. 


from  the  ends  of  the  impeller  to  the  discharge  outlet.  The  impellers 
may  be  of  the  open  type,  Fig.  291(B),  or  closed,  Fig.  291(A).  The  casing 
may  be  cylindrical  and  concentric  with  the  impeller,  Fig.  295,  or  of 


PUMPS 


561 


spiral  form,  Fig.  292.  The  shape  of  the  impeller  and  casing  determines 
the  efficiency  of  the  pump  and  its'  adaptability  to  certain  conditions  of 
service. 

Centrifugal  pumps  may  be  classified  as 

(1)  Volute. 

(2)  Turbine. 

293.  Volute  Pumps.  —  Fig.  292  shows  a  section  through  a  typical 
volute  pump.     The  casing  is  of  spiral  design,   forming  a  gradually 

increasing  water  or  "  whirlpool  "  chamber 
for  the  purpose  of  partially  converting 
velocity  head  to  pressure  head.  The  older 
forms  of  volute  pumps  were  very  ineffi- 
cient, seldom  delivering  more  than  50  per 
cent  of  the  energy  supplied,  and  usually 
not  adapted  to  lifts  greater  than  50  feet. 
The  modern  pump  gives  efficiencies  as 
high  as  77  per  cent,  and  lift  is  limited  only 
by  the  speed  of  the  impeller.  As  a  general 
rule  the  volute  pump  is  of  single-stage 
construction,  and  best  adapted  to  com- 
paratively low  lifts  (under  80  feet),  though 
an  exception  is  found  in  the  De  Laval 

steam  turbine  driven  volute  pumps,  which  are  made  both  single  and 

multi-stage  for  lifts  as  high  as  700  feet. 

294.  Turbine  Pumps.  —  The  directions  of  flow  in  the  casing  and  from 
the  impeller  in  a  volute  pump  are  at  cross  currents  with  each  other,  as 
shown  in  Fig.  293.     The  turbine  pump,  Fig.  294,  is  provided  with  a 


FIG.    292. 


SUCTION 

A  Typical  Centrifugal 
Pump. 


FIG.  293.    Direction  of  Water  from  the  Impellers 
of  a  Centrifugal  Pump  without  Diffusion  Vanes. 


FIG.  294.     Effect  of  Diffusion  Vanes 
on  the  Direction  of  Water. 


system  of  diffusion  vanes  or  expanding  ducts  disposed  between  the 
periphery  of  the  impeller  and  the  annular  casing  somewhat  like  the  guide 
vanes  in  a  reaction  turbine  water  wheel,  so  that  the  fluid  emerges 


562 


STEAM  POWER  PLANT  ENGINEERING 


tangentially  at  about  the  velocity  in  the  casing.     The  casing  is  con- 
centric with  the  impeller,  since  the  diffusion  vanes  render  the  volute 


FIG.  295.     Lea-Degan  Three-Stage  Turbine  Pump. 

chamber  unnecessary.  For  high  lifts  these  pumps  are  compounded, 
thereby  reducing  the  peripheral  velocity  and  decreasing  the  friction 
losses.  Fig.  295  shows  a  section  through  a  three-stage  turbine  pump 
and  Fig.  296  a  section  through  a  six-stage  pump. 


FIG.  296.     Rateau  Six-Stage  Turbine  Pump. 

In  view  of  past  developments  it  is  probable  that  the  centrifugal  pump 
will  supplant  to  a  considerable  extent  the  present  type  of  piston  pump 


PUMPS  563 

for  many  uses.  Efficiencies  above  70  per  cent  are  not  unusual,  and 
the  head  against  which  the  pump  may  operate  is  limited  only  by  the 
peripheral  speed  at  which  the  impeller  may  be  safely  run.  Some  of 
the  advantages  of  the  modern  high-grade  centrifugal  pump  as  compared 
with  the  piston  type  are: 

(1)  Low  first  cost, 

(2)  Compactness, 

(3)  Absence  of  valves  and  pistons, 

(4)  Low  rate  of  depreciation, 

(5)  Uniform  pressure  and  flow  of  water, 

(6)  Simplicity  of  design  and  ease  of  operation, 

(7)  Freedom  from  shock, 

(8)  High  rotative  speed,  permitting  direct  connection  to  electric 

motors  and  steam  turbines, 

(9)  Ability  to  handle  dirty  water,  sewage,  and  the  like, 

(10)  In  case  of  stoppage  of  delivery,  the  pressure  cannot  increase 

beyond  the  predetermined  working  pressure,  and 

(11)  Ease  of  repair. 

Some  of  the  disadvantages  are: 

(1)  Efficiency  not  as  high  as  the  best  grade  of  piston  pumps, 

(2)  Cannot  be  direct  connected  to  low-speed  engines  when  high 

lifts  are  desired,  and 

(3)  The  rate  of  flow  cannot  be  efficiently  regulated  for  wide  ranges 

in  duty. 

Theory  of  Centrifugal  Pumps:  Engr.  U.S.,  Oct.  1,  1907,  p.  908;  Bulletin  of  Univ. 
of  Wisconsin,  No.  173;  Trans.  A.S.M.E.,  54-470;  Prac.  Engr.,  Lond.,  Sept.  29,  1905; 
Jour.  Asso.  Eng.  Soc.,  Dec.,  1901 ;  Centrifugal  Pumps,  Turbines,  and  Water  Motors, 
C.  H.  Innes,  Chap.  XXVI. 

Centrifugal  Pumps,  General  Description:  Trans.  A.S.M.E.,  26-764;  Engr.  U.S., 
Aug.  1,  1907,  p.  723,  Oct.,  1907,  p.  908,  Oct.  15,  1907,  p.  952;  Power,  March,  1907, 
p.  172;  Eng.  Mag.,  July,  1906;  Elec.  World,  Jan.  12,  1907,  p.  113;  Eng.  Rec.,  Feb.  9, 
1907,  p.  165;  Machinery,  Jan.,  1907,  p.  237,  April,  1907,  p.  442. 

295.  Performance  of  Centrifugal  Pumps.  —  For  best  efficiency  a 
centrifugal  pump  must  be  properly  designed  for  the  intended  service 
as  to  curvature  of  vanes,  diameter  and  speed  of  impeller,  and  num- 
ber of  stages.  Figs.  298  to  300  are  based  upon  experiments  with 
De  Laval  centrifugal  pumps.  When  a  practically  uniform  head  is 
required  at  constant  speed  with  varying  water  supply  as  in  city  water- 
works, hydraulic  elevator  systems,  or  boiler  feeding,  the  impeller  vanes 
are  designed  to  give  the  characteristic  curve  illustrated  in  Fig.  298, 
which  protects  the  motor  from  possible  overload. 


564 


STEAM  POWER  PLANT  ENGINEERING 


In  dry-dock  and  other  variable-head  work,  in  order  not  to  overload 
the  motor,  the  power  should  be  practically  constant  through  wide 
variations  of  head  and  at  the  same  time  the  efficiency  should  not  vary 
seriously.  A  desirable  characteristic  for  such  a  pump  is  illustrated  in 
Fig.  299. 


50 


Curves  of  Performance 

For 
Morris  Ceutrifugal  Pump 

Variable  Speed 

Bated  Cap.  For  60  Ft.  Hd. 

120  Gal.  at  830  R.P.M. 


400  600 

Revolutions  Per  Min. 

FIG.  297. 


In  water-supply  systems  in  which  the  friction  of  the  piping  is  a  large 
part  of  the  total  head  at  full  delivery,  the  characteristic  shown  in  Fig. 
300  is  especially  useful.  Thus,  when  the  system  reduces  its  demand 
for  water  and  the  frictional  head  is  consequently  considerably  reduced, 


160 
140 
120 

ioo 

80 
60 
40 


-we 


20 


40 


60 


120 


140 


FIG. 


80  100 

Capacity 

Centrifugal    Pump   Characteristic    for   Hydraulic    Elevator    Service,    Boiler 
Feeding,  etc. 


20  40  60  80  100  120  140 

Capacity 
FIG.  299.    Centrifugal  Pump  Characteristic  for  Dry-Dock  Service. 


160 


140 


120 


100 


Chara 


U   i 


20 


40 


60  80  100  120  140 

Capacity 
FIG.  300.    Centrifugal  Pump  Characteristic  for  Water  Works  with  Large  Friction  Head. 


566 


STEAM  POWER  PLANT  ENGINEERING 


£ 
I 

•        i-J 

CO 

1s-        ^3 

O 

3  E 

«    I 


G 


«PI  jo  jooj 
qrag;  JQJ 


utj\[  jad  ' 


•UJK 


•sgqouj 


•saqouj 
'uoipng 


•saqDUj 


^CO(Mi-i 


eOC^COOOr-  ( 


1-1  T-H  1-1  <M  -^  CO  O> 


PUMPS 


567 


the  pump  would  automatically  adjust  itself  to  the  reduced  head  without 
change  of  speed.  Figs.  297  to  304  are  based  upon  experiment  and  show 
the  relationship  between  speed;  head,  capacity,  efficiency,  and  power 
consumption  of  various  types  of  pumps. 

Tables  73  and  74  give  the  capacity,  speed,  head,  and  power  require- 
ments for  commercial  sizes  of  centrifugal  pumps,  and  may  be  used  as  a 
guide  in  selecting  the  size  of  pump  for  general  service. 

TABLE   74. 
DATA  PERTAINING  TO  WORTHINGTON  MULTI-STAGE  TURBINE  PUMPS. 


~ 

, 

Total  Head  in  Feet,  R.P.M.,  Number  of  Stages. 

1 

3 

3:| 

£ 

*S  •'* 

!  j 

£| 

|§ 

100  Ft. 

200  Ft. 

300  Ft. 

400  Ft. 

0  {I, 

0    g 

Cil       ^ 

l-o 

o    . 

0      . 

•3  w. 

•g 

1    f  1 

Is 

1  8 

^ 

Ui     0) 

i>   be 

jjj 

L-  <§ 

<u   So 

g- 

1  s 

g- 

Oi    §x) 

o3    0  i-t 

|s 

"S  J> 

a  fe 

fc 

|l 

PL; 

C   rr\ 
3 

PH' 

ll 

PM 

§  02 
9 

5 

s 

3 

* 

ti 

* 

f4 

P4 

ti 

5 

1  .5 

30 

0.02 

2000 

2 

1.5 

2 

45-60 

0.0395 

1500 

2 

2066 

3 

2 

2  5 

75-100 

0.0625 

1300 

2 

1800 

3 

1500 

4 

2.5 

3' 

125-150 

0.095 

1200 

2 

1600 

3 

1300 

4 

5 

3 

4 

200-250 

0.134 

1100 

2 

1400 

3 

1200 

4 

5 

4 

5 

350-450 

0.222 

950 

2 

1200 

3 

1000 

4 

5 

5 

6 

600-700 

0.297 

800 

1 

1300 

2 

1150 

3 

1050 

4 

6 

7 

800-1000 

0.396 

750 

1 

1200 

2 

1000 

3 

950 

4 

8 

10 

1500-1800 

0.643 

600 

1 

1000 

2 

800 

3 

780 

4 

10 

12 

2500-2800 

1.00 

500 

1 

800 

2 

700 

3 

670 

4 

*  Horse  power  based  on  maximum  capacity. 

Tests  of  Centrifugal  Pumps:  Engr.  U.S.,  Oct.  15,  1906,  p.  685;  Eng.  and  Min.  Jour., 
April  14,  1906,  p.  698;  Eng.  News,  June  2,  1904,  p.  512;  Eng.  Rec.,  July  1,  1905, 
p.  25,  Sept.  29,  1906,  p.  352;  Iron  Age,  Sept.  1,  1904,  p.  25;  Machinery,  Nov.,  1906, 
p.  144;  Power,  Nov.,  1906,  p.  688;  Trans.  A.S.M.E.,  22-262,  831;  Jour.  Am.  Soc. 
Naval  Engrs.,  17-85. 

296.  Rotary  Pumps.  —  Rotary  pumps  are  often  used  for  circulating 
cooling  water  in  condenser  installations,  and  give  about  the  same 
efficiency  as  centrifugal  pumps  under  similar  conditions  of  operation. 
For  moderate  pressure  and  large  volumes  they  offer  the  advantage  of 
low  rotative  speed,  thus  permitting  direct  connection  to  slow-speed 
steam  engines.  At  high  speeds  they  are  noisy,  due  chiefly  to  the 
gearing.  They  occupy  considerably  less  space  than  piston  pumps  of 
the  same  capacity,  but  require  more  room  than  the  centrifugal  type. 

Fig.  305  shows  a  section  through  a  two-lobe  cycloidal  pump.  The 
shafts  are  connected  by  wheel  gearing,  the  power  being  applied  to  one 
of  the  shafts.  The  water  is  drawn  in  at  /  and  forced  out  at  0,  the 


568 


STEAM  POWER  PLANT  ENGINEERING 


in 

[-j 

40 

He 

ac 

— 

^^^ 

-^* 

9 

-  — 

^, 

; 

. 

^. 

30 

.^ 

^ 

s 

^ 

^ 

\ 

" 

•j 

1 

u 

\ 

•s 

/ 

? 

V 

/ 

/. 

/ 

/ 

/ 

/ 

/ 

/ 

1 

/ 

«U 

11 

10 

& 

10 

HI 

10 

41 

X) 

5C 

K> 

dc 

0 

7( 

X) 

8( 

X) 

a 

X) 

10G 

Capacity,  Gallons  per  Minute. 
FIG.  301.     Performance  of  a  Six-Inch  Worthington  Conoidal  Pump. 


.SINGLt  STAGE  PUMP,  DESIGNED  FOR 
1700  OALL.  P.M.  *  100  FEET  AT  1545  R.P. 


400  eoo  eoo  /oo  eoo  BOO  <ooo  noo  1200  1300  >*oo  1000 


Performance  of  a  Single-Stage  De  Laval  Centrifugal  Pump. 


PUMPS 


569 


TOTAL  Lirr.  FCE 


e2 


20 


GALLONS  PER  MINUTE 


555 555 — 600    eoo   1000  izoo  woo  ieoo  1000  eooo  aaoo  a4<x>  asoo  zaoo  3000  3zoo  3400  aeoo  3000 
FIG.  303.     Performance  of  a  Two-Stage  Lea-Degan  Turbine  Pump. 


1100 

PUMP  oesiGN£o  FOR  zso  GALL.  P.M.  x  ••foo  FEET 

*T  200O  AND  20000  R.P.M. 
TE6TIO  APRIL  1TM  AND  *TH,  1«04 

1 

f 

2 

1 

• 
£ 

TOT4 

C£fr 

2 

• 
• 
K 

eoo 

700 

eoo 

500 
400 
JOO 
300 
100 

» 

1 
i 

-4 

I 

^ 

\ 

I 

0 

5 

\ 

* 

'^ 

\ 

\ 

1 

5 

*i» 
N^ 

• 

v 

w 

\ 

| 

p^l 
* 

&t 

» 

-r 

^ 

3 
3 

4 

*,  * 

1 

5 

4 

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V-tj 

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v 

i 

i 

/ 

%-. 

t^c/ 

1 

0 

1 

f 

OALLC 

US  PER  ' 

I.NUTE 

s 

l-l 

0  M>  100          ISO          200          JSQ 

FIG.  304.    Performance  of  a  Two-Stage  De  Laval  Centrifugal  Pump. 


570 


STEAM  POWER  PLANT  ENGINEERING 


FIG.  305.    Two-Lobe  Cycloidal 
Pump. 


FIG.  306.     Rotary    Pump  with 
Movable  Butment. 


50 


70 


N 

1  65 


2.8 

2.6 
2.4 

§  2J2 

1  2.0 

OH 

2  1.8 

S,e 

1.4 


1.2 


400 


500 


600  700  800 

Revolutions  Per  Minute 


900 


Head  Constant,  Speed  Variable. 


o^: 


1000 


Average  Speed  812  R.P.M. 
Average  Capacity 
45.5  Gal.Per  Minute 


70         80          90         100       110        120        130        140       150 
Total  Head/ Eeet, 

Speed  Constant,  Head  Variable. 


170     180 


307.    Performance  of  a  Small  Rotary  Pump. 


PUMPS 


571 


displacement  per  revolution  being  equal  to  four  times  the  volume  of 
chamber  A.  There  is  no  rubbing  between  impellers  and  casing.  In 
this  type  of  pump  the 
pressure  is  independent 
of  the  speed  of  rotation, 
and  the  capacity  varies 
almost  directly  with  the 
speed.  The  slip  varies 
from  5  to  20  per  cent 
according  to  the  dis- 
charge pressure. 

Fig.  306  shows  a  sec- 
tion through  a  rotary 
pump  with  movable  but- 
ment.  Fig.  307  illus- 
trates the  performance 
of  a  45-mm.  Siemens- 
Schuckert  rotary  pump 
at  different  speeds  and 
discharge  pressures. 
(Zeit.  d.  Ver.  Deut.  Ing., 
June  24,  1905,  p.  1040.) 
Large  rotary  pumps  give 
much  higher  efficiencies, 
but  the  general  charac- 
teristics are  about  the 
same.  A  combined  effi- 
ciency of  pump  and 
engine  as  high  as  84  per 
cent  has  been  recorded. 
(Trans.  A.S.M.E.,  Vol. 
24,  p.  385.) 

Rotary  Pumps:  Engr. 
U.S.,  Jan.  1,  1904,  p.  51; 
Am.  Mach.,  March  12,  1896, 
p.  238,  Jan.  25,  1906,  p.  103  ; 
Trans.  A.S.M.E.,  24-385, 
Eng.  News,  March  1,  1900; 
p.  152;  Power,  Dec.,  1903, 

p.   59,  Aug.,    1905,    p.    477;  FIG.  308.    10,000,000  Gallon  Circulating  Pt^rip" 

The  Constructor,  Releaux,  p.  226;  Eng.  Rec.,  Jan.  10,  1903,  p.  59. 

Tests  of  Rotary  Pumps:   Zeit.  d.  Ver.  Deut.  Ing.,  June  24,  1905,  p.  1040;  Trans. 
A.S.M.E.,  28—503. 


572 


STEAM   POWER   PLANT  ENGINEERING 


297.  Circulating  Pumps.  —  This  term  is  ordinarily  applied  to  the 
pumps  which  supply  injection  water  to  the  condenser. 

These  types  are  found  in  practice:  the  piston,  the  centrifugal,  and 
the  rotary  pump.  Figs.  224  and  236  show  the  application  of  recipro- 
cating pumps  to  condenser  installations  and  Fig.  21  la  and  Fig.  443 
a  similar  application  of  centrifugal  pumps. 

For  large  volumes  of  water  and  low  heads  the  centrifugal  or  rotary 

pump  is  generally  adopted  on  ac- 
count of  minimum  space  require- 
ments and  low  first  cost. 

In  very  large  central  stations 
where  the  demand  for  circulating 
water  is  enormous  and  the  lift  is 
moderately  high,  the  high-duty 
pumping  engine  is  often  installed. 
Fig.  308  shows  a  section  through 
one  of  the  nine  high-duty  circulat- 
ing pumps  at  the  New  York  Rapid 
Transit  Company's  power  house. 
The  steam  end  is  operated  by 
Corliss  cylinders  and  is  of  the 
cross  compound  type.  The  maxi- 
mum capacity  is  10,000,000  gallons 
per  day  (24  hours)  against  a  head 
of  50  feet  at  mean  low  water.  The 
actual  lift  is  much  less  than  this, 
as  the  discharge  is  aided  by  the 
vacuum  in  the  condenser. 

Pumping  Engines:  Engr.  U.S.,  Dec.  1, 
1905,  p.  786;  Engng.,  Jan.  27,  1905,  p.  132; 
Eng.  Mag.,  Jan.,  1905, p-  77 ;  Eng.  Rec.,  July 
8,  1905,  p.  58;  Eng.  News,  Dec.  6,  1905, 
May  26,  1904;  Trans.  A.S.M.E.,  3-141, 
9-476,  12,  534,  975,  13,  83,  176,  14-1340, 
FIG.  309.  The  Pulsometer.  15-1103,  16,  49,  169,  21,  327,  788,  1018. 

298.  Air  Lift.  —  The  air  lift  is  a  simple  arrangement  of  piping 
whereby  water  may  be  raised  by  means  of  compressed  air.  There  are 
no  working  parts,  and  no  valves  are  employed  except  to  regulate  the 
supply  of  air.  Its  particular  field  of  application  lies  in  pumping  water 
from  a  number  of  scattered  wells,  and  on  account  of  the  total  absence 
of  working  parts  it  is  peculiarly  adapted  to  handling  water  containing 
sand,  grit,  and  the  like  The  device  consists  of  a  partially  submerged 


PUMPS 


573 


water  pipe  and  an  air-supply  pipe  variously  arranged  as  in  Fig.  310(A) 
to  310(D).  Compressed  air  forced  into  the  water  pipe  at  or  near  the 
bottom  forms  a  series  of  bubbles  or  "  pistons/'  as  shown  in  D,  which 
displace  an  equal  volume  of  water.  (For  the  theory  of  the  air  lift  see 
Compressed  Air,  October,  1905,  p.  3696.)  The  pressure  required  to 
operate  the  lift  after  it  is  once  started  is  considerably  less  than  the 


WATER   LEVEL 


FIG.  310.     Different  Arrangements  of  the  "  Air  Lift." 

pressure  of  water  due  to  the  head,  while  that  required  to  start  it  is 
slightly  greater,  consequently  the  pressure  may  be  reduced  after  the 
pump  begins  to  work  properly.  The  successful  operation  of  this 
device  depends  upon  the  ratio  of  the  depth  of  submersion  A,  Fig. 
310(D),  to  the  total  lift  B,  and  the  ratio  of  the  area  of  the  air  pipe  to 
that  of  the  water  pipe.  The  best  results  are  obtained,  in  a  general 

sense,  when  the  ratio  —  lies  between  0.55  and  0.85,  an  average  figure 
n 

being  0.65,  and  when  the  area  of  the  air  pipe  is  0.16  that  of  the  water 
pipe. 

The  quantity  of  air  needed  may  be  closely  approximated  by  the 
following  equation  (Engineer,  London,  Aug.  14,  1903,  p.  173): 


20 
in  which 

V  =  cubic  feet  of  free  air  per  minute. 

Q  =  cubic  feet  of  water  per  minute. 

L  =  lift  in  feet  above  the  surface  of  the  water. 

The  velocity  of  the  air  should  not  exceed  4000  feet  per  minute. 


574 


STEAM  POWER  PLANT  ENGINEERING 


The    efficiency    ("  water"    horse    power    divided    by  "air"    horse 

^ 

power)    varies   from  30   to   50   per   cent,   increasing   as   the   ratio  -r 

B 

increases  from  0.55  to  0.85.  (Engineer,  U.S.,  Aug.  15,  1904,  p.  564.) 
A  number  of  tests  give  efficiencies  ("  water  "  horse  power  divided  by 
I.H.P. -of  steam  cylinder)  varying  from  20  to  40  per  cent.  The  horse 
power  required  to  compress  one  cubic  foot  of  free  air  to  different 
pressures  per  square  inch,  as  determined  from  actual  practice,  is 
approximately  as  follows: 


Horse  Power 

Horse  Power 

Pressure  in 

Required  to 

Pressure  in 

Required  to 

Pounds. 

Compress 

Pounds. 

Compress 

1  Cubic  Foot. 

1  Cubic  Foot. 

176 

0.434 

60 

0.159 

140 

0.376 

45 

0.145 

100 

0.201 

30 

0.121 

80 

0.189 

(Engr.,  Lond.,  Aug.  14,  1903,  p.  174,  Dec.  11,  1903,  p.  568,  Feb.  12,  1904,  p.  172.) 

When  it  becomes  necessary  to  raise  water  to  a  height  exceeding 
say  175  feet  above  the  level  in  the  well,  it  is  customary  to  use  two  or 
more  pumps,  the  total  lift  being  divided  between  them. 

Air  Lift:  Engr.,  Lond.,  Aug.  14,  1903,  p.  173,  Dec.  11,  1903,  p.  568,  Feb.  12,  1904, 
p.  172,  Jan.  10,  1908;  Eng.  Rec.,  Jan.  7,  1905,  p.  8;  Engineering,  Jan.  22,  1904, 
p.  135,  Jan.  29,  1904,  p.  166;  Compressed  Air,  Aug.,  1903,  Oct.,  1905;  Engr.  U.S.,  July 
15,  1903,  p.  547,  Sept.  21,  1906,  Aug.  15,  1904,  p.  564;  Mech.  Engr.,  Aug.  20,  1904; 
Engng.  Rev.,  July,  1904;  Power,  March,  1905,  p.  173;  Eng.  News,  Jan.  16,  1908. 

Pulsometer:  Tech.  Quar.,  Sept.,  1901;  Public  Works,  Aug.  15,  1904;  Engr.  U.S., 
July  15,  1904;  Experimental  Eng.,  Carpenter,  p.  621;  Thermodynamics,  Wood, 
p.  293;  Trans.  A.S.M.E.,  13-211. 

Cost  of  Operating  American  Pumping  Stations:  Eng.  Rec.,  Aug.  6,  1904;  Proc. 
Engrs.  Club  of  Phil.,  Oct.,  1906. 

Complete  Description  of  Various  American  Types  of  Steam,  Rotary,  and  Centrifugal 
Pumps:  Engr.  U.S.,  Jan.  1,  1904. 

The  Growth  of  the  Pumping  Station:   Eng.  Rec.,  July  14,  1906,  p.  50. 

The  Selection  of  Waterworks  Pumping  Machinery:   Eng.  News,  Vol.  52,  p.  39. 

Centrifugal  Pump  for  Boiler  Feeding:  Power  &  Engr.,  Mar.  15,  1910. 

Recent  Records  of  High  Duty  Pumping  Engine:  Eng.  News,  Feb.  3,  1910. 


CHAPTER  XIV. 

SEPARATORS,  TRAPS,  DRAINS. 

299.  Live-Steam  Separators.  General.  —  The  function  of  a  steam 
separator  is  the  removal  of  entrained  water  from  steam. 

Unless  a  boiler  is  liberally  provided  with  superheating  surface,  the 
steam  may  contain  an  amount  of  moisture  varying  from  0.3  to  5  per 
cent.  If  the  boiler  is  poorly  proportioned  or  forced  far  above  its  rating, 
this  percentage  may  be  greatly  increased.  The  quality  of  the  steam 
is  still  further  reduced  by  condensation  in  the  steam  pipe,  which  may 
vary  from  one  to  ten  per  cent,  depending  upon  the  length  of  pipe  and 
efficiency  of  covering. 

One  of  the  effects  of  moisture  in  steam  is  to  increase  its  density 
and  reduce  its  elastic  force.  It  also  increases  its  conductivity,  so  that 
during  the.  work  of  expansion  more  heat  is  absorbed  from  the  walls  of 
the  cylinder  and  discharged  into  the  atmosphere  or  into  the  condenser 
without  doing  useful  work.  (Ewing,  "The  Steam  Engine,"  p.  151.) 
Although  the  heat  loss  from  this  cause  is  small,  the  danger  arising 
from  the  introduction  of  a  considerable  amount  of  water  in  the 
cylinder  renders  the  removal  of  the  moisture  necessary.  See  page  248 
for  influence  of  moisture  on  steam  consumption. 

The  essentials  of  a  good  separator  are  high  efficiency  as  a  water 
eliminator,  ample  storage  capacity  for  any  sudden  influx  of  water, 
simplicity  and  durability  in  construction,  and  small  resistance  to  the 
current  of  steam  passing  through.  A  good  separator  may  be  relied 
upon  to  remove  practically  all  of  the  moisture  from  steam  containing 
under  ten  per  cent  entrainment  and  all  but  two  per  cent  from  steam 
containing  as  much  as  twenty  per  cent.  (Engineer,  U.S.,  Jan.  15, 
1904.) 

Table  75  gives  the  results  of  a  series  of  tests  made  by  Professor 
R.  C.  Carpenter  in  1891  of  six  steam  separators.  (Power,  July,  1891, 
p.  9.)  Conclusions  from  these  tests  were: 

1.  That   no   relation   existed   between  the  volume   of  the  several 
separators  and  their  efficiency. 

2.  No  marked  decrease  in  pressure  was  shown  by  any  of  the  separa- 
tors, the  most  being  1.7  pounds  by  separator  E. 

575 


576 


STEAM  POWER  PLANT  ENGINEERING 


3.  Although  changed  direction,  reduced  velocity,  and  perhaps  centri- 
fugal force  are  necessary  for  good  separation,  still  some  means  must 
be  provided  to  lead  the  water  out  of  the  current  of  the  steam. 

A  series  of  tests  made  at  Armour  Institute  of  Technology  in  1905  on 
a  number  of  separators  showed  that  the  efficiency  of  separation  decreased 
as  the  velocity  of  the  steam  increased.*  At  the  low  velocity  of  500  feet 
per  minute  all  separators  were  equally  efficient,  at  a  velocity  of  5000 
feet  per  minute  several  had  little  effect  on  eliminating  the  moisture 
present,  and  at  a  velocity  of  8000  feet  per  minute  only  one  gave  efficient 
results. 

TABLE   75. 
TESTS  OF  STEAM  SEPARATORS. 

(R.  C.  Carpenter.) 


Test  with  Steam  of  about  10 
per  Cent  of  Moisture. 

Tests  with  Varying  Moisture. 

Make  of 

Separator. 

Quality  of 
Steam 

Quality  of 
Steam 

Efficiency. 

Quality  of 
Steam  Before. 

Quality  of 
Steam  After. 

Average 
Efficiency. 

Before. 

After. 

Per  Cent. 

Per  Cent. 

Per  Cent. 

Per  Cent. 

Per  Cent. 

Per  Cent. 

B 

87  0 

98  8 

90  8 

66  1-97  5 

97  8-99 

87  6 

A 

90  1 

98  0 

80  0 

51  9-98 

97  g_99  i 

76  4 

D         .     . 

89  6 

95  8 

59  6 

72  2-96  1 

95  5-98  2 

71  7 

c  

90  6 

93  7 

33  0 

67  1-96  8 

93  7-98  4 

63  4 

E  

88.4 

90.2 

15.5 

68.6-98.1 

79.3-98.5 

36.9 

F  

88.9 

92.1 

28.8 

70.4-97.7 

84.1-97.9 

28.4 

300.   Classification  of  Separators.  —  Separators  are  based  on  one  or 
more  of  the  following  principles  of  action: 

1.  Reverse  current.     The  direction  of  the  flow  is  abruptly  changed, 
usually  through  180  degrees.     This  causes  the  water  in  the  steam,  on 
account  of  its  greater  specific  gravity,  to  be  thrown  into  a  receiving 
vessel,  while  the  steam  passes  on  in  a  reverse  direction. 

2.  Centrifugal  force.     A  rotary  motion  is  imparted   to   the   steam 
whereby  entrained  water  particles  are  eliminated  by  centrifugal  force. 

3.  Baffle  plates.     The  flow  is  interrupted  by  corrugated  or  fluted 
plates  to  the  surfaces  of  which  the  water  particles  adhere  and  from 
which  they  fall  by  gravity  to  the  well  below. 

4.  Mesh.     The  separation  is  brought  about  by  mechanical  filtration 
through  screens  or  meshes. 

The  following  outline  shows  the  classification  of  typical  separators, 
in  accordance  with  the  above  principles : 

*  See  Power,  May  11,  1909,  p.  834. 


SEPARATORS,  TRAPS,  DRAINS 


577 


Live-steam  separators 


Reverse  current 


Centrifugal 


Baffle  plate 


Mesh 


Exhaust-steam  separators 


( Jacketed  baffle  . . 


Hoppes,  Fig.  311. 
Stratton,  Fig.  312. 
Keystone,  Fig.  313. 
Mosher. 
Robertson. 
Bundy,  Fig.  314. 
Austin,  Fig.  315. 
Detroit. 

Direct,  Fig.  316. 
Potter  mesh. 
.Baurn,  Fig.  317. 


(Absorption Loew,  Fig.  318. 


301.  Reverse-Current  Steam  Separators.  —  Fig.  311  shows  a  section 
through  a  Hoppes  steam  separator  and  illustrates  the  principle  of 
reverse-current  separation.  Steam  may 
flow  through  in  either  direction.  Both 
the  inlet  and  outlet  ports  are  surrounded 
by  gutters  C,  C,  partly  filled  with  water, 
which  intercept  the  moisture  following 
the  surface  of  the  pipe,  while  the  down- 
ward plunge  of  the  steam  throws  the 
entrained  water  to  the  bottom  of  the 
separator.  The  condensation  is  carried 
from  the  troughs  by  pipe  P  to  the  well 
below,  from  which  it  is  trapped  at  D  in 
the  usual  way.  The  velocity  of  the 
steam  in  passing  through  this  separator 
is  greatly  reduced  to  prevent  the  steam 
from  taking  up  the  water  in  the  bottom 
of  the  well.  This  is  brought  about  by 
increasing  the  area  of  the  passage  through 
the  separator. 

Fig.  312  gives  a  sectional  view  of  a  Stratton  separator,  which,  though 
primarily  of  the  reverse-current  type,  embodies  also  the  principle  of 
centrifugal  force.  The  separator  consists  of  a  vertical  cast-iron  cylinder 
with  an  internal  central  pipe  C  extending  from  the  top  downward 
for  about  half  the  height  of  the  apparatus,  leaving  an  annular  space 
between  the  two.  The  current  of  steam  on  entering  is  deflected  by  a 
curved  partition  and  thrown  tangentially  to  the  annular  space  at  the 
side,  near  to  top  of  the  apparatus.  It  is  thus  whirled  around  with  all 
the  velocity  of  influx,  producing  the  centrifugal  action  which  throws 
the  particles  of  water  against  the  outer  cylinder.  These  adhere  to  the 
surface,  so  that  the  water  runs  down  continuously  in  a  thin  sheet  around 


FIG.  311.     Hoppes  Steam  Sepa- 
rator. 


578 


STEAM  POWER  PLANT  ENGINEERING 


the  outer  shell  into  the  receptacle  below.  The  steam,  following  in  a 
spiral  course  to  the  bottom  of  the  internal  pipe,  abruptly  enters  it,  and 
passes  upward  and  out  of  the  separator  without  having  once  crossed 
the  stream  of  separated  water.  The  rapid  rotation  of  the  current  of 
steam  imparts  a  whirling  motion  to  the  separated  water  which  tends  to 
interfere  with  its  proper  discharge  from  the  apparatus.  The  separator 
has  therefore  been  provided  with  wings  or  ribs  E  projecting  at  an  acute 
angle  to  the  course  of  the  current,  which  have  the  effect  of  breaking  up 
this  whirling  motion  and  allowing  the  water  to  settle  quietly  at  the 
bottom,  whence  it  passes  off  through  the  drain  pipe  D. 


FIG.  312.    Stratton  Steam  Separator. 


FIG.  313.     Keystone  Steam  Separator. 


302.  Centrifugal  Steam  Separators.  —  Fig.  313  shows  a  section 
through  a  Keystone  or  Simpson's  centrifugal  separator.  The  separator 
consists  of  a  cast-iron  cylinder  with  vertical  pipe  C  extending  down- 
ward about  two-thirds  of  the  whole  length;  this  pipe  has  a  thread  or 
screw  wound  spirally  around  it,  the  space  between  the  threads  being 
somewhat  greater  than  the  area  of  the  steam  pipe.  The  steam  passing 
around  the  spiral  course  causes  the  water  to  be  thrown  against  the 
outer  walls  by  centrifugal  force,  while  the  dry  steam  passes  through 


SEPARATORS,  TRAPS,  DRAINS 


579 


FIG.  314.     Bundy 
Steam  Separator. 


the  small  holes  in  the  central  pipe.  The  water  passes  down  the  outer 
walls,  where  its  motion  is  arrested  by  obstructing  ribs  E,  and  is  thence 
carried  away  by  a  drip  pipe  D  to  a  suitable  drain. 

303.  Baffle-Plate   Steam    Separators.  —  Fig.    314   gives    an   interior 
view  of  a  Bundy  separator  and  illustrates  the  application  of  baffle 
plates  for  live  steam  separation.      This  separator 

consists  of  a  rectangular  cast-iron  casing  with  a 
cylindrical  receiver  beneath  it.  Directly  across  the 
steam  passage  are  baffle  plates  corrugated  for  the 
reception  of  entrained  water.  The  plates  consist 
of  vertical  castings,  each  containing  a  main  artery 
or  channel  which  leads  directly  to  the  receiver. 
The  fronts  of  the  plates  are  flat,  with  a  series  of 
recesses  sloping  inwards  and  downwards,  terminat- 
ing in  an  opening  of  capillary  size  leading  to  the 

main  artery.  The  plates 
are  staggered,  so  that 
the  steam  must  impinge 
against  all  of  them  in  its 
passage.  The  particles 
of  water  adhere  to  the  plates,  collect,  and 
fall  by  gravity  into  the  receiver.  The 
flanges  at  the  bottom  constrict  the  opening 
of  the  reservoir  so  as  to  prevent  the  steam 
from  picking  up  any  portion  of  the  water. 
Fig.  315  shows  a  section  through  an 
Austin  separator  and  illustrates  another 
class  embodying  the  fluted  baffle  plate 
principle.  The  steam  in  passing  through 
the  chamber  impinges  against  the  fluted 
baffle,  plate  B.  The  moisture  adheres  to 
the  surfaces,  collects,  and  trickles  along  the 
corrugations  to  the  bottom  of  the  well. 
These  corrugations  are  formed  in  such  a 
manner  that  the  steam  cannot  come  in 
contact  with  the  water  particles  after 
they  have  been  once  eliminated.  A  per- 
forated diaphragm  D  prevents  the  water  in  the  well  from  coming 
in  contact  with  the  steam.  The  current  of  steam  is  also  reversed, 
thus  giving  additional  separating  properties  to  the  apparatus. 

304.  Mesh  Separators.  —  Fig.  316  shows  a  section  through  a  "  direct  " 
separator,  illustrating  the  principle  of  mesh  separation.     These  separa- 


FIG.  315. 


Austin  Steam  Sepa- 
rator. 


580 


STEAM   POWER  PLANT  ENGINEERING 


tors  are  made  with  steel  bodies  and  cast-iron  heads-  and  bases,  in  all 
sizes  up  to  six  inches  inclusive,  the  larger  sizes  being  constructed  of 
cast  iron  or  boiler  plate.  The  cone  C,  perforated  lining  E,  and  dia- 
phragm S  are  made  of  cold-rolled  copper;  the  cone  0  is  a  substantial 

gray-iron  casting,  resting  on  three  cast- 
iron  supports  hooked  over  the  top  of 
inner  pipe  as  indicated.  The  method 
of  operation  is  as  follows :  The  accumu- 
lated moisture  around  the  walls  of  the 
steam  pipe  is  caught  by  the  upper  edge 
of  cone  C  and  carried  down  back  of 
lining  E  to  *  the  water  chamber.  The 
current  of  steam  entering  the  separator 
impinges  upon  the  conical  surface,  which 
is  composed  of  solid  plate  0  covered 
with  sieve  S,  through  which  water  may 
freely  pass  but  from  which  it  cannot 
readily  escape.  Passing  through  the 
sieve  and  depositing  on  the  solid  surface 
of  the  cone  0,  this  water  is  carried  by 
conductors  P  to  the  water  chamber. 
Perforated  lining  E  permits  the  moisture 
content  of  the  steam  to  pass  through  the 
opening  to  the  water  below  and  prevents 
it  from  coming  in  contact  again  with  the 
current  of  steam.  A  trough  is  provided 
at  the  lower  edge  of  the  inverted  cup 
which  leads  all  the  water  that  may  ad- 
here to  it  to  the  water  chamber.  The 

steam  flows  through  the  passages  indicated  by  arrows  and  is  sub- 
jected to  a  whipsnapping  action  which  tends  to  throw  off  any  re- 
maining moisture.  The  perforated  plate  D  prevents  the  steam  from 
picking  water  out  of  the  water  chamber. 

305.   Location.  —  Live-steam  separators  may  be  located 

1.  Inside  the  boiler. 

2.  Between  boiler  and  engine. 

3.  At  the  steam  chest. 

Where  the  steam  pipe  is  very  short,  and  particularly  in  marine  and 
locomotive  work  where  the  tossing  of  the  boiler  induces  excessive 
priming,  the  separator  may  be  placed  inside  the  boiler  and  its 
function  becomes  that  of  a  dry  pipe.  In  this  location  it  prevents  the 
water  due  to  foaming  and  priming  from  passing  to  the  engine,  and 


H 


FIG.   316.     "  Direct  "  Steam    Sepa- 
rator. 


SEPARATORS,  TRAPS,  DRAINS  581 

reduces  condensation  in  the  pipe  by  supplying  dry  steam.  The 
"  Potter  mesh  "  and  the  "  De  Rycke  centrifugal  "  are  types  of  sepa- 
rators designed  for  this  service. 

The  arrangement  of  separator  between  engine  and  boiler,  other  than 
at  the  throttle,  or  inside  the  boiler  is  sometimes  necessary  for  economy 
of  space.  Where  possible,  however,  the  separator  should  be  placed 
close  to  the  steam  chest. 

Current  practice  recommends  that  a  receiver  separator,  which  is  an 
ordinary  separator  with  a  volume  of  two  to  four  times  that  of  the 
high-pressure  cylinder,  be  placed  close  to  the  engine  if  the  load  is 
intermittent  or  sharply  fluctuating.  This  forms  a  cushion  for  absorb- 
ing the  force  of  the  blows  caused  by  cut-off,  delivers  steam  at  a  prac- 
tically uniform  pressure,  and  reduces  the  vibration  of  the  piping  to  a 
minimum.  It  also  provides  a  reservoir  for  sudden  demands  made  by 
the  engine.  Smaller  pipes  and  higher  velocities  may  be  used  with  this 
arrangement. 

306.  Exhaust-Steam  Separators  and  Oil  Eliminators.  —  The  function 
of  an  exhaust-steam  separator  is  the  removal  of  cylinder  oil  from  the 
steam  exhausted  by  engines  and  pumps.  In  plants  where  exhaust 
steam  is  used  for  heating  it  is  quite  essential  to  remove  the  oil  from 
the  steam  before  it  enters  the  heating  system,  for  the  oil  not  only 
reduces  the  efficiency  of  the  radiators  by  coating  them  with  an  excel- 
lent non-conducting  film  but  is  an  element  of  danger  to  the  boiler 
itself.  In  condensing  plants  the  separator  will  prevent  the  oil  from 
fouling  the  condenser  tubes  and  those  of  the  vacuum  heater  if  one  is 
installed;  this  is  an  important  factor,  since  the  oil  or  grease  lowers  the 
efficiency  of  the  heat  transmission. 

In  a  general  sense  a  live-steam  separator  is  also  an  oil  eliminator,  and 
all  the  separators  previously  described  perform  this  function  to  a  cer- 
tain extent,  since  the  underlying  principles  governing  the  elimination 
of  oil  from  exhaust  steam  are  similar  to  those  employed  in  removing 
water  from  steam.  Most  of  the  separators  described  above  are  also 
designed,  in  lighter  form,  as  oil  eliminators,  but  by  far  the  greater 
number  are  based  on  the  fluted  baffle  plate  principle,  of  which  the  Hine, 
Bundy,  Cochrane,  Utility,  Peerless,  and  Keiley  are  well-known  examples. 
This  type  of  oil  separator  will  eliminate  a  considerable  portion  of  the 
oil  in  the  steam  provided  the  baffle  plates  or  corrugated  surfaces  are 
frequently  cleaned. 

The  following  is  taken  from  the  report  of  Professor  R.  Burnham  of 
the  Armour  Institute  of  Technology  on  the  test  of  a  six-inch  horizontal 
oil  separator  of  the  baffle-plate  type: 

"  For  purposes  of  test  the  separator  was  placed  in  the  exhaust  line 


582 


STEAM  POWER  PLANT  ENGINEERING 


of  a  9  x  18  x  24  cross  compound  Corliss  engine  running  under  its  maximum 
load  at  80  pounds  pressure  and  exhausting  into  a  Wheeler  surface 
condenser  against  26  inches  vacuum. 

"Cylinder  oil  was  fed  through  the  lubricators  of  the  high  and  low 
pressure  cylinders  at  the  rate  of  from  5  to  20  drops  per  minute,  a  record 
being  made  of  the  exact  quantity  of  oil  fed  per  hour.  The  separator 
was  so  arranged,  by  means  of  a  receiver  connected  to  the  air  pump,  that 
the  accumulation  of  oil  and  water  could  be  readily  trapped  from  it  at 
any  time.  In  order  to  determine  the  quantity  of  oil  given  up  by  the 
condenser,  and  not  properly  charged  against  the  separator,  each  series 
of  efficiency  tests  was  preceded  by  a  run  of  three  hours  during  which 
time  no  oil  whatever  was  fed  to  the  cylinders.  During  the  last  hour  a 
record  was  made  of  the  weight  of  steam  used  and  a  sample  of  the  con- 
denser discharge  retained  for  analysis. 

"  The  efficiency  tests  were  made  by  feeding  at  an  excessive  rate 
through  the  lubricators  as  described  above,  and  when  conditions  became 
practically  constant,  records  were  made  for  one  hour,  of  the  weight  of 
oil  used,  weight  of  condensed  steam,  and  drain  from  separator.  Samples 
of  the  two  latter  were  retained  for  analysis  and  the  percentage  of  oil  in 
them  accurately  determined,  correction  being  made  for  the  oil  given  up 
by  the  condenser.  A  second  series  of  tests  was  made  exhausting  at 
atmospheric  pressure.  The  results  obtained  are  tabulated  below. 


Exhausting  into 
26-inch  Vacuum. 

Exhausting  at 
Atmospheric 
Pressure. 

Oil  in  condensed  steam  with  no  oil  feeding. 
(Charged  to  condenser.)    Pounds  per  hour 
Oil  fed  to  cylinder,  pounds  per  hour  
Steam  condensed  per  hour,  pounds  
Oil  caught  by  separator,  per  hour,  pounds  A 
Oil  in  condensed  steam  (corrected),  pounds 
per  hour              B 

.051     .057   .0559 
.401     .562   .934 
1000     1120  1096 
.341     .450  .743 

009       010     0096 

.0353       .0340 
.621          .710 
-    905          872 
.552         .583 

0071         0050 

Percentage  of  oil  in  condensed  steam  by 
weight  per  cent                                        •  •  • 

ftflftQ     fl01      00088 

00078       00057 

Q7    A      Q7   8      Q8    8 

Q8   7          QQ    1 

"  There  was  practically  no  free  oil  on  the  surface  of  the  condenser 
discharge  in  any  case,  the  small  quantity  of  oil  which  passed  the  separa- 
tor (from  5  to  10  parts  in  a  million  of  water  by  weight)  existing  as  an 
emulsion,  imparting  a  slight  milky  color  to  the  water." 

It  is  a  well-established  fact  that  oil  can  be  more  effectually  removed 
from  wet  than  from  dry  steam,  and  some  makers,  notably  the  Austin 


SEPARATORS,  TRAPS,  DRAINS 


583 


FROM  ENGINE 


Separator  Company,  inject  a  cold-water  spray  into  the  separator  cham- 
ber. A  similar  result  is  brought  about  in  the  Baum  separator,  Fig.  317, 
in  which  the  corrugated  baffle  plate  is  hollow  and  cold  water  is  forced 
through  the  chamber  thus  formed.  Referring  to  Fig.  317:  The  diverged 
baffle  plate  forms  the  wall  of  a  chamber  in  which  cold  water  is  con- 
tinually circulated.  This  circulation 
causes  moisture  to  appear  on  the  baffle 
plate  surface.  The  particles  of  oil, 
coming  in  contact  with  this  moist  sur- 
face as  the  steam  current  is  diverged, 
adhere  to  it  and  fall  by  gravity  into  the 
well  below,  where  they  are  completely 
isolated  from  the  purified  steam.  A  large 
portion  of  the  oil  and  water,  however, 
does  not  enter  the  separator  at  all  but 
is  caught  by  the  inside  ledge  near  the 
junction  of  the  exhaust  pipe  and  the 
separator.  The  oil  and  condensation 
which  are  carried  along  the  bottom  of 
the  pipe  come  in  contact  with  thisjedge 
and  are  carried  directly  to  the  outlet 
pipe. 

A  very  successful  method  of  removing 
oil  from  steam  is  to  project  the  steam  on 
to  the  surface  of  a  body  of  water.  The 
water  may  be  hot  or  cold  and  will  hold 
the  oil  if  it  once  reaches  the  surface.  It 
is  essential,  however,  to  reduce  the  velo- 
city of  the  steam  as  it  passes  on  its  way 
to  the  outlet.  Baldwin's  grease  separator  is  based  upon  this  principle. 
(Baldwin  on  Heating,  p.  234.) 

The  most  efficient  method  of  removing  oil  is  by  combined  filtration 
and  absorption.  (Engineering  News,  May  22,  1902,  p.  406.)  A  large 
chamber  filled  with  coke,  brick,  broken  tile,  or  other  absorption  material 
is  placed  in  series  with  the  exhaust  pipe.  The  steam  passing  through 
this  chamber  is  entirely  freed  from  oil  and  moisture,  provided  the  absorb- 
ing material  is  sufficient  in  quantity  and  is  replenished  as  soon  as  it 
becomes  saturated  with  oil.  The  annoyance  attending  the  removal 
and  replenishing  of  the  absorbing  material  at  frequent  intervals  and 
the  great  size  of  the  apparatus  are  serious  drawbacks.  An  example  of 
this  system  of  purification  in  which  many  of  the  objectionable  features 
are  reduced  to  a  minimum  is  the  Loew  grease  and  oil  separator,  Fig.  318. 


DRAIN 

FIG.  317.     Baum  Oil  Separator. 


584 


STEAM  POWER  PLANT  ENGINEERING 


The  exhaust  steam  enters  the  chamber  at  the  top,  strikes  a  large 
deflecting  plate  shaped  like  an  inverted  V,  and  permits  part  of  the 
condensation  and  oil  to  be  drawn  off  by  the  drain  pipe.  The  steam 
then  rises  and  is  deflected,  as  indicated,  against  a  series  of  shelves 
filled  with  fibrous  material  covered  with  coarse  wire  screens.  The 
grease  is  removed  from  each  shelf  by  suitable  drains.  This  apparatus 

is  sectional  and  any  number  of  sections 
may  be  added  without  affecting  the 
rest. 

In  a  non-condensing  plant  where  the 
exhaust  steam  is  used  for  heating  pur- 
poses the  oil  separator  is  ordinarily 
placed  in  the  main  exhaust  pipe  just 
before  it  enters  the  heating  system. 
Where  several  branches  enter  one  main 
it  is  not  customary  to  place  a  sepa- 
rator in  each  branch,  one  large  separator 
located  as  above  being  sufficient. 

In  condensing  plants  oil  separators 
are  seldom  installed  except  where  sur- 
face condensers  are  used,  in  which 
case  the  separator  may  be  placed 
anywhere  between  the  engine  and 
condenser.  In  case  a  vacuum  heater 
„  _  is  used  the  separator  may  be  placed 

FIG.  318.    Loew  Grease  Extractor.  ... 

on  either  side  of  the  heater,  depend- 
ing upon  the  type  of  separator.  If  the  separator  is  of  the  "  jacket 
cooling  "  or  "  spray  "  type,  it  may  be  placed  between  the  engine  and 
the  vacuum  heater;  if.  however,  it  is  of  the  "  baffle-plate"  type,  the  oil 
will  be  more  efficiently  removed  if  the  separator  is  placed  between 
the  heater  and  condenser  so  that  it  will  get  the  benefit  of  the 
moisture  formed  in  the  heater.  In  the  latter  location,  however, 
the  separator  will  not  prevent  the  oil  from  fouling  the  heater 
tubes. 

Where  a  jet  condenser  is  used  and  water  is  taken  from  the  hot  well, 
the  hot  well  itself  acts  as  an  oil  separator.  (Trans.  A.S.M.E.,  24- 
1144.) 

All  separators,  steam  and  oil,  should  be  provided  with  gauge  glasses 
and  should  be  thoroughly  drained  and  the  drainage  should  be 
automatic. 

307.  Exhaust  Heads.  —  The  function  of  the  exhaust  head  is  the 
elimination  of  oil  and  water  from  steam  exhaust  before  permitting  it 


SEPARATORS,  TRAPS,  DRAINS 


585 


to  be  discharged  into  the  atmosphere.  Unless  removed,  the  water 
and  oil  rot  the  roofs  and  walls  in  summer  and  pollute  the  atmosphere 
surrounding  the  plant.  The  exhaust  head  also  acts  as  a  muffler 
reducing  the  noise  of  the  escaping  steam.  Exhaust  heads  are  built 

on  the  same  principle  as  steam  and 
oil  separators  and  most  separator 
builders  manufacture  them.  Fig. 
319  shows  a  section  through  a  typ- 
ical exhaust  head.  The  condensa- 
tion is  ordinarily  drained  to  waste, 
though  with  proper  purification  it 
may  be  returned  to  the  boiler. 
With  an  efficient  oil  separator  in 
the  exhaust  line  the  condensation 
in  the  exhaust  head  may  be  returned 
directly  to  the  boiler  without  further 
purification. 

Live-steam  separators  are  propor- 
tioned so  that  it  is  only  necessary,  in 
the  average  installation,  to  specify 
the  size  of  pipe,  the  type  of  engine, 
the  steam  pressure,  and  the  style, 
whether  horizontal  or  vertical.  Gauge 
glasses,  gauge  cocks,  and  companion 
flanges  are  usually  provided  by  the 
maker.  In  some  cases  the  capacity 
of  the  reservoir  is  also  specified.  In 


FIG.  319.    A  Typical  Exhaust-Head. 


specifying  oil  extractors  the  following  additional  data  are  necessary 
for  an  intelligent  choice :  the  number  of  engines  and  pumps  exhausting 
into  the  line,  the  location  of  the  separator,  the  steam  pressure,  velocity, 
and  the  quality  and  quantity  of  cylinder  oil  used.  A  guarantee  of 
efficiency  and  of  material  and  workmanship  is  often  demanded. 


REFERENCES. 

Water  and  Oil  Separators:   Am.  Elecn.,  Jan.,  1900. 

Steam  Separators:   Am.  Elecn.,  June,  1905. 

"  Dry  Steam  ":    Goubert  Mfg.  Co.,  85  Liberty  St.,  N.Y.  (Catalogue). 

Cochrane  Separator:  Harrison  Safety  Boiler  Works,  Philadelphia,  Pa. 
(Catalogue). 

Bundy  Separator:   A.  A.  Griffin  Iron  Co.,  N.Y.  (Catalogue). 

A  Bad  Case  of  Discharge  Water  with  Steam  from  Water-Tube  Boilers:  Trans. 
A.S.M.E.,  Vol.  26. 

Location  of  Steam  Separators:   Power,  Oct.,  1904. 


586  STEAM  POWER  PLANT  ENGINEERING 

Condensing  Exhaust  Head:  Eng.  News,  Vol.  49,  p.  419;  Eng.  Rec.,  Vol.  40, 
p.  177. 

Experiments  in  Separating  Oil  from  Condensed  Steam:  Eng.  News,  May  22,  1903; 
Eng.  Rec.,  April  27,  1901;  Engr.,  Lond.,  March  12,  1897,  Oct.  20,  1905;  Power, 
Aug.,  Sept.,  1896,  May,  1903;  Heating  and  Ventilation,  Feb.,  1897;  Trans.  A.S.M.E. 
Vol.  17,  p.  295. 

Oil  Separators:   Baldwin  on  Heating,  pp.  233-237. 

Oil  Separation  in  a  Combination  Engine  and  Turbine  Plant:   Power,  Oct.,  1906. 

Test  of  a  Cochrane  Steam  Separator:   Power,  April,  1898. 

Test  ofLippincott  Separator:   Engr.  U.S.,  Aug.,  1902. 

Test  of  a  Linstrum  Steam  Separator:   Engr.  U.S.,  June  15,  1904. 

Test  of  an  Austin  Steam  and  Oil  Separator:   Trans.  A.S.M.E.,  Vol.  20,  p.  489. 

Test  of  a  Detroit  Live  Steam  Separator:  Engr.  U.S.,  April  15,  1904;  Power,  Jan., 
1902. 

Tests  of  Direct  Separators,  Oil  and  Steam:  Direct  Separator  Co.,  Syracuse,  N.Y. 
(Catalogue). 

Tests  of  Six  Steam  Separators:  Power,  July,  1891,  p.  9;  Engr.  News,  Vol.  26, 
p.  233. 

Test  of  a  "  Utility  "  Oil  Eliminator:  Engr.  Rec.,  May  2,  1903;  Engr.  Rev.,  May, 
1903. 

The  Hot  Well  as  an  Oil  Extractor:  Trans.  A.S.M.E.,  24-1144. 

308.  Drips.  —  No  matter  how  thoroughly  a  steam  pipe  or  reservoir 
may  be  covered  with  insulating  material  considerable  condensation 
takes  place.     With  the  best  covering  this  loss  approximates  one-sixth 
of  a  pound  of  steam  per  square  foot  of  pipe  surface  per  hour  for  steam 
pressures  of  one  hundred  pounds,  and  runs  as  high  as  one  pound  of 
steam  for  bare  pipes.     See  Table  80  for  results  of  experiments  on  the 
loss  of  heat  from  bare  pipes,  and  Table  81  for  data  on  the  efficiency 
of  pipe  coverings.     In  addition  to  this  water  of  condensation,  from  \ 
to  2  per  cent  of  moisture  is  carried  over  by  the  steam  from  the  boiler. 
This  water,  unless  thoroughly  removed,  is  a  constant  source  of  danger 
to  the  engines  and  causes  water  hammer  and  leaky  joints  in  the  piping. 

A  joint  on  a  steam  pipe  may  safely  withstand  a  steam  pressure  of 
100  pounds  without  leaking  and  still  leak  badly  under  a  water  pressure 
of  half  that  amount.  This  is  due  to  the  fact  that  the  steam  with  its 
high  temperature  causes  the  pipe  to  expand,  thus  insuring  a  tight 
joint,  while  the  entrained  water  (which  cools  as  it  collects)  causes  the 
pipe  to  contract  and  allows  a  leak. 

The  entrained  water  and  water  of  condensation  are  usually  spoken  of 
as  "  drips."  Drips  may  be  divided  into  two  classes,  low  pressure  and 
high  pressure. 

309.  Low-Pressure  Drips.  —  Low-pressure  drips  include  the  steam 
condensed  in  exhaust  steam  feed  heaters  of  the  closed  type,  exhaust 
steam  piping,  receiver  barrels,  steam  chests,  and  exhaust  heads.     As 


SEPARATORS,  TRAPS,  DRAINS 


587 


these  drips  are  impregnated  with  oil  and  are  useless  for  boiler  feed 
without  purification,  they  are  usually  discharged  to  waste.     Most  city 


FIG.  320.     Closed  Heater  Installation  for  Abstracting  Heat  from  Oily  Drips. 

ordinances  require   the  drips   to  be  cooled  to  100   degrees  F.  before 

being  discharged  into  the  sewer.  In  this  case  they  must  be  first  dis- 
charged into  a  tank  and  permitted  to 
cool.  This  tank  must  be  vented  to  the 
atmosphere  to  prevent  back  pressure. 
Fig.  320  shows  an  installation  in  which 
the  heat  asbtr acted  from  the  drips,  etc., 
is  used  to  heat  the  feed  water.  The 
drips  from  the  throttle  valve  and  steam 
chest  in  a  non-condensing  plant  are 
ordinarily  discharged  into  the  exhaust 
pipe  as  shown  in  Fig.  321.  In  a  con- 
densing plant  the  throttle  drips  are 
piped  to  a  trap  or  to  the  free  exhaust 
pipe.  The  returns  from  a  steam- 
heating  system  are  sometimes  classi- 

FIG.  321.    Simple  Method  of  Draining  fied   as    low-pressure  drips.     They   are 

Dnps*  invariably  returned  to  the  boiler. 

In  small  plants  all  the  low-pressure  drips  may  be  connected  to  one 

large  pipe  and  this  pipe  in  turn  to  a  single  trap,  provided  there  is  but 


588  STEAM  POWER  PLANT  ENGINEERING 

little  difference  in  pressure  in  the  various  drip  pipes.    In  case  of  different 
pressures  separate  leads  should  be  run  to  waste  or  traps. 

The  drips  from  the  receiver  and  vacuum  heater  barrels  in  a  con- 
densing plant  are  oftentimes  under  less  than  atmospheric  pressure,  and 
sometimes  the  pressure  varies  from  a  slight  vacuum  to  10  or  20  pounds 
gauge,  and  consequently  cannot  be  disposed  of  as  described  above. 
If  possible,  the  heaters  and  receivers  should  be  placed  so  as  to  drain  into 
the  condenser  (see  Fig.  334).  Should  this  arrangement  prove  impracti- 
cable, the  barrels  may  be  drained  by  a  trap  especially  arranged  as  shown 
in  Fig.  335. 

310.  Size  of  Pipe  for  Low-Pressure  Drips.  —  In  the  average  exhaust- 
steam  feed-water  heater  one  pound  of  steam  in  condensing  gives  up 
approximately  1000  heat  units.     This  will  heat  about  6  pounds  of  water 
from  60  to  200  degrees  F.     Hence  the  area  of  the  drip  which  carries  the 
water  of  condensation  from  the  closed  heater  need  be  but  J  that  of  the 
feed  pipe.    In  no  case,  however,  should  a  pipe  smaller  than  J  inch  in  diam- 
eter be  used.     Should  the  same  pipe  be  used  for  both  exhaust  head  and 
heater  drips,  an  area  of  J  area  of  feed  pipe  would  prove  of  ample  capacity. 
In  practice  it  is  customary  to  use  the  size  of  pipe  conforming  with  the 
outlet  furnished  by  the  manufacturer  of  the  apparatus,  and  only  when 
several  pieces  of  apparatus  are  connected  to  one  main  are  calculations 
made  for  the  size  of  this  main. 

The  drip  pipe  from  the  throttle  valve  is  ordinarily  J  inch  diameter 
irrespective  of  the  size  of  steam  pipe;  this  is  also  true  of  the  steam-chest 
drip. 

311.  High-Pressure  Drips.  —  High-pressure  drips  consist    of    those 
which  are  condensed  under  practically  boiler  pressure  and  include  the 
steam  condensed  in  steam  pipes,  cylinder  jackets  of  engines,  reheating 
coils  of  receivers,  and  separators.     Being  free  from  oil  and  containing 
considerable  heat,  they  are  usually  returned  to  the  boiler.     Drips  may 
be  returned  to  the  boiler  automatically  by  means  of 

1.  Steam  traps. 

2.  Holly  steam  loop. 

3.  Pumps. 

312.  Classification  of  Steam  Traps.  —  Steam  traps  may  be  'divided 
into  two   classes,   depending  on  their  use,  —  return  and  non-return. 
Both  of  these  two  classes  may  be  subdivided  into  five  types  according 
to  the  principle  of  operation,  viz. : 

I.    Float.  III.    Bowl. 

II.   Bucket.  IV.    Expansion. 

V.    Differential. 


SEPARATORS,  TRAPS,  DRAINS 


589 


Return  Traps. 

Traps  which  receive  the  condensed  steam  and  return  it  to  a  boiler 
having  considerably  higher  pressure  than  that  acting  on  the  returns 
are  known  as  return  traps.  They  are  made  in  a  great  variety  of  styles. 
The  general  principle  of  operation  is  shown  in  Fig.  331  and  described  in 
paragraph  318. 

Non-Return  Traps. 

Non-return  traps,  as  the  name  implies,  are  used  where  the  water  of 
condensation  is  not  returned  to  the  boiler  but  is  discharged  into  any 
receptacle  having  less  than  boiler  pressure. 

CLASSIFICATION  OF  A  FEW  WELL-KNOWN  STEAM  TRAPS. 


Steam  Traps , 


fFloat  .  . 

[McDaniel. 

Bucket  .  . 

;  Cookson. 
\  Acme. 

• 
Dump  . 

\  Albany. 
^Bundy. 

" 

Morehead. 
fMetal  

(  Columbia. 

Expansion  
Differential  .      ... 

Volatile-Fluid  .  . 
Flinn. 

(  Geipel. 
(  Dunham. 
(  Heintz. 

Siphon. 

FIG.  322.     McDaniel  Float  Trap. 

313.  Float  Traps.  —  Fig.  322  shows  a  section  through  a  McDaniel 
improved  trap,  illustrating  the  principles  of  the  float  type.  A  hollow 
sphere  C  of  seamless  copper  pivoted  at  E  rises  and  falls  with  the  change 
of  water  level  in  the  vessel.  The  discharge  valve  M  is  operated  by  the 


590 


STEAM  POWER  PLANT  ENGINEERING 


float.  When  the  trap  is  empty  the  float  is  in  its  lowest  position  and 
the  discharge  valve  is  closed.  Water  of  condensation  flows  into  the 
trap  by  gravity  through  opening  D  to  a  certain  depth,  when  the  float 
opens  the  discharge  valve  and  the  steam  pressure  acting  on  the 
surface  of  the  water  forces  it  through  outlet  S  to  tank  or  atmos- 
phere. After  the  water  is  discharged  the  float  closes  the  valve  and 
permits  the  condensation  to  collect  again.  A  gauge  glass  indicates  the 
height  of  water  in  the  chamber. 

Unless  float  traps  are  well  made  and  proportioned  there  is  a  danger 
of  considerable  steam  leakage  through  the  discharge  valve,  due  to 
unequal  expansion  of  valve  and  seat  and  the  sticking  of  moving  parts. 
The  discharge  from  a  float  trap  is  usually  continuous,  since  the  height  of 
the  float,  and  consequently  the  area  of  the  outlet,  is  proportional  to  the 
amount  of  water  present.  When  the  trap  is  working  lightly,  this 
adjustment  is  apt  to  throttle  the  area  and  create  such  a  high  velocity  of 
discharge  as  to  cause  a  rapid  wear  of  valve  and  seat.  This  defect  is 
more  or  less  evident  in  all  steam  traps  discharging  continuously.  For 
this  reason  all  wearing  parts  should  be  accessible  and  readily  replaced. 


FIG.  323.     Acme  Bucket  Trap. 

314.  Bucket  Traps.  —  Fig.  323  shows  a  section  through  an  "  Improved 
Acme  "  steam  trap.  The  water  of  condensation  enters  the  cast-iron 
vessel  at  A,  filling  the  space  D  between  the  bucket  E  and  the  walls  of 


SEPARATORS,  TRAPS,  DRAINS 


591 


the  trap.  This  causes  the  bucket  to  float  and  forces  valve  V  against 
its  seat  (valve  V  and  its  stem  being  fastened  to  the  bucket  as  indicated). 
When  the  water  rises  above  the  edges  of  the  bucket  it  flows  into  it  and 
causes  it  to  sink,  thereby  withdrawing  valve  V  from  its  seat.  This 
permits  the  steam  pressure  acting  on  the  surface  of  the  water  in  the 
bucket  to  force  the  water  through  the  annular  space  H  to  discharge 
opening  G.  When  the  bucket  is  emptied  it  rises  and  closes  valve  V 
and  another  cycle  begins.  By  closing  valve  R  the  trap  is  by-passed 
and  the  condensation  blows  directly  through  passage  C  to  discharge  G. 
The  discharge  from  this  type  of  trap  is  intermittent. 

315.   Dump  or  Bowl  Traps.  —  Fig.   324  shows  sections   through   a 
Bundy  bowl  trap  of  the  "  return  "  type.     The  water  enters  the  bowl 


FIG.  324.     A  Typical  Tilting  Trap. 


through  trunnion  D  and  rises  until  its  weight  overbalances  counter- 
weight E  and  the  bowl  sinks  to  the  bottom.  As  the  bowl  sinks,  arm 
G,  which  is  a  part  of  the  bowl,  rises  and  engages  the  nuts  N  on  valve 
stem  H  and  opens  valve  7,  thus  admitting  live  steam  pressure  on  to  the 
surface  of  the  water.  The  trap  then  discharges  like  all  others.  After 
the  water  is  discharged  weight  E  sinks  and  raises  bowl  A,  which  in  turn 
closes  valve  /,  and  the  cycle  begins  again.  Bowl  traps  are  necessarily 
intermittent  in  their  discharge. 

Fig.  335  shows  the  application  of  a  bowl  trap  to  a  receiver  where  the 
drips  are  under  a  vacuum,  and  Fig.  336  a  similar  application  to  an 
engine  receiver  where  the  pressure  varies  from  less  than  atmospheric 
pressure  to  a  pressure  of  40  or  50  pounds. 


592  STEAM  POWER   PLANT  ENGINEERING 

316.  Expansion  Traps.  —  Expansion  traps  may  be  divided  into  two 
groups: 

(1)  Those  in  which  the  discharge  valve  is  operated  by  the  relative 
expansion  of  metals  and 

(2)  Those  in  which  the  action  of  a  volatile  fluid  is  utilized. 
Expansion  traps  will  never  freeze,  as  they  are  open  when  cold  and 

all  the  water  drains  out  before  the  freezing  temperature  is  reached. 

Since  traps  of  this  type  have  little  capacity  for  holding  water,  5  to 
10  feet  of  pipe  should  be  provided  between  the  trap  and  the  pipe  to  be 
drained  in  order  that  the  condensation  may  collect  and  cool. 

Fig.  325  shows  the  general  appearance  of  a  Columbia  expansion 
trap  in  which  the  valve  is  operated  by  the  expansion  of  metallic  tubes. 


INLET 


OUTLET 

FIG.  325.    A  Typical  Expansion  Trap. 

Water  gravitates  to  the  trap  through  opening  marked  "  inlet/'  passes 
through  brass  pipe  0,  then  downward  to  the  main  body  of  the  valves 
and  back  to  outlet  valve  C.  Below  pipe  0  and  parallel  to  it  is  an 
iron  rod  S,  at  the  end  of  which  is  the  support  or  fulcrum  of  lever  R. 
The  lower  end  of  this  lever  is  connected  to  the  stem  of  the  valve  C,  so 
that  any  movement  of  the  lever  is  communicated  to  it.  When  the  trap 
is  cold,  valve  C  is  open  and  all  water  of  condensation  passes  out.  The 
moment  steam  enters  the  pipe  0  it  expands.  The  amount  of  expansion 
is  multiplied  several  times  by  the  action  of  the  lever  R,  so  that  the 
movement  of  the  valve  is  much  greater  than  the  expansion  of  the 
pipe  0.  The  compensating  spring  D  prevents  the  brass  tube  from 
damaging  itself  by  excessive  expansion.  Lever  A  permits  the  trap 
to  be  blown  through  by  hand. 

Fig.  326  shows  a  section  through  a  Geipel  trap  in  which  the  valve 
is  operated  directly  by  the  expansion  of  two  metallic  tubes  and  the 
movement  is  not  multiplied  by  levers  as  with  the  Columbia.  The 
lower  or  brass  pipe  constitutes  the  inlet,  and  is  connected  to  the  vessel 
to  be  drained;  the  upper  or  iron  pipe  is  the  outlet  for  discharge.  The 
two  pipes  form  the  sides  of  an  isosceles  triangle,  the  base  F  of  which  is 


SEPARATORS,  TRAPS,  DRAINS 


593 


rigid*  while  the  apex  A  is  free  to  move  in  a  direction  at  right  angles  to 
the  linear  expansion  of  the  tubes.  When  cold,  the  brass  pipe  is  con- 
tracted and  the  apex,  in  which  the  valve  seat  is  placed,  is  moved  down 
so  that  the  valve  is  open  and  the  water  is  discharged.  As  soon  as  steam 


FIG.  326.     Geipel  Expansion  Trap. 

enters  the  brass  pipe  the  latter  expands  and  forces  the  valve  seat 
against  the  valve.  The  trap  may  be  adjusted  for  any  pressure  by 
means  of  the  lock  nuts  E.  When  it  is  desired  to  blow  through,  the 
valve  may  be  operated  by  hand  by  pressing  the  lever. 

Fig.  327  shows  a  section  through  a  Dunham  trap.     It  operates  upon 
the  expansion  principle,  utilizing  a  fluid  of  a  volatile  character  as  its 


FIG.  327.     Dunham  Expansion  Trap. 

motive  force.  The  corrugated  bronze  disk  B  is  filled  with  a  volatile 
fluid,  and  expands  and  contracts  according  to  the  pressure  exerted  by 
the  fluid.  The  water  enters  at  the  top,  surrounds  disk  B  and  passes 
through  valve  opening  D  to  discharge  outlet  at  E.  As  soon  as  steam 
strikes  the  disk  B  the  volatile  fluid  flashes  into  a  vapor  and  causes  the 
disk  to  expand.  This  expansion  forces  valve  D  against  its  seat  and 
the  discharge  ceases.  The  valve  will  remain  closed  until  the  con- 
densation collects  and  cools  the  disk  B,  which  then  contracts,  opens 


594 


STEAM  POWER  PLANT  ENGINEERING 


the  valve,  and  condensation  enters  as  before.  The  adjustment  f  how- 
ever, is  such  that  the  discharge  may  be  made  continuous  instead  of 
intermittent. 

The  Dunham  trap  is  claimed  to  be  the  smallest  trap  of  its  capacity 
on  the  market.  The  1-inch  size,  having  a  capacity  for  draining 
10,000  lineal  feet  of  1-inch  pipe  under  60  pounds  pressure,  weighs 
but  5  pounds  and  may  be  connected  to  the  pipe  Una  as  if  it  were  a 
globe  valve. 

Fig.  328  shows  an  internal  view  of  a  Heintz  steam  trap.  This 
works  on  the  principle  of  the  volatile  fluid  expansion  trap  but  in 


Fig.  328.     Heintz  Expansion  Trap. 

a  different  manner  from  any  of  those  described  above.  The  requisite 
movement  is  obtained  by  the  elongation  and  contraction  of  the 
extremities  of  a  bent  metallic  tube  T  filled  with  a  highly  volatile 
fluid.  This  tube  is  inclosed  in  a  cast-iron  box  and  presses  against 
the  point  of  regulating  screw  P.  The  other  extremity  of  the  tube 
carries  the  valve  and  is  free  to  move  under  the  action  of  the 
variations  of  temperature.  Spring  S  has  no  connection  with  the 
action  of  the  trap.  It  is  used  as  a  simple  means  of  holding  one 
end  of  the  expansion  tube  on  its  pivot.  The  trap  operates  as  follows: 
Water  enters  at  /,  surrounds  the  tube  T  and  passes  through  the 
valve  to  the  discharge  outlet  O.  As  soon  as  steam  enters  the  chamber 
the  volatile  fluid  in  the  tube  flashes  into  a  vapor  and  the  pressure 
thus  created  tends  to  straighten  out  the  tube;  this  forces  the  valve 
against  its  seat  and  the  discharge  ceases.  As  the  trap  cools  the  tube 
returns  to  its  normal  position  and  the  discharge  valve  is  opened,  thus 
permitting  the  condensation  to  drain  out.  The  adjustment  permits  of 
continuous  or  intermittent  discharge  and  of  variable  pressures. 

317.  Differential  Traps.  —  Fig.  329  shows  a  cross  section  through 
a  Flinn  differential  trap.  The  column  of  water  X  acting  on  dia- 
phragm D  closes  valve  V.  The  water  entering  pipe  E  and  the 
action  of  the  spring  equalize  column  X  and  open  the  valve. 


SEPARATORS,  TRAPS,  DRAINS 


595 


Describing  the  action  in  further  detail,  the  water  of  condensation 
enters  at  A,  fills  lower  chamber  Y,  pipe  X,  and  receiving  chamber  C 
up  to  the  level  of  the  top  of  pipe  E.  This  column  of  water  acting  on 
the  under  side  of  the  diaphragm  D  forces  the  valve  to  its  seat  against 
the  counter  pressure  of  the  spring  S.  Any  additional  water  that 
enters  the  trap  overflows  through  pipe  E,  filling  chamber  F  and  pipe  E 
to  a  point  about  midway  of  its  height,  where  the  effect  of  the 
column  of  water  in  pipe  X  is  balanced.  The  pressure  on  each  side  of 
the  diaphragm  is  then  equal,  the  short  column  in  pipe  E,  aided  by 
the  spring,  balancing  the  pressure  of  the  longer  column  in  pipe  X. 

Any  further  increase  in  the  height 
of  the  water  in  pipe  E  causes  a 
depression  of  the  valve  V,  which 
allows  water  to  escape  until  the 
column  has  fallen  to  a  level  a 
little  below  the  middle  of  pipe  E, 
when  this  valve  closes  again.  This 
action  is  repeated  at  intervals 
according  to  the  quantity  of  water 
entering  the  trap. 
So  long  as  the  water 
keeps  coming  in  ~ 
sufficiently  large 
quantities  the  valve 
remains  wide  open. 
Fig.  330  gives  a 
general  view  of  a 
siphon  trap  which 
is  much  used  in 
draining  low-pres- 
sure systems,  as, 
for  example,  the 
separator  in  an  ex- 
haust steam  heat- 
ing system.  It  con-  F'G:3U30-  SimPle 

.,  .*••„          ,        Siphon  Trap. 

sists  essentially  of 
two  legs  A  and  B,  which  may  be  close  together  or  any  distance  apart 
but  the  lengths  of  which  must  be  sufficiently  great  to  prevent  pressure 
acting  through  pipe  7  from  forcing  the  water  out  of  B.  C  is  a  vent  pipe 
extending  to  the  air  to  prevent  siphoning;  0  is  the  discharge  for  the 
condensed  steam.  In  ordinary  operation  the  leg  B  is  filled  with  water 
which  is  constantly  overflowing,  and  A  with  steam  and  water,  the 


Flinn  Differential  Trap. 


596 


STEAM  POWER  PLANT  ENGINEERING 


total  pressure  in  both  legs  being  equal.  The  siphon  trap  is  applicable 
for  low  pressure  only,  as  it  requires  approximately  2.3  feet  of  vertical 
space  E  for  each  pound  per  square  inch  pressure  in  the  pipe.  The 
maximum  allowable  head  is  represented  by  vertical  distance  N. 

318.  Location  of  Traps.  —  Wherever  possible  a  trap  should  be 
located  so  that  the  condensation  will  flow  into  it  by  gravity.  This 
will  insure  positive  drainage.  Sometimes,  however,  the  coils,  cylin- 
ders, or  pipes  to  be  drained  are  located  in  a  pit  or  trench  or  lie  on  a 
basement  floor  where  it  is  impossible  to  set  the  trap  so  as  to  receive 
the  drains  by  gravity  without  placing  it  in  an  inaccessible  position. 
With  very  low  pressures  this  is  often  unavoidable,  but  with  pressures  of 
five  pounds  or  more  the  trap  may  be  placed  above  the  point  to  be 
drained.  If  a  trap  is  set  in  an  exposed  place  a  drain  should  be  pro- 
vided at  the  lowest  point  to  free  the  pipe  of  water  when  steam  is  shut 
off.  A  dirt  catcher  or  strainer  should  be  placed  in  the  pipe  leading  to 
the  trap  to  prevent  scale,  etc.,  from  reaching  the  valve.  All  pockets 
and  dead  ends  should  be  drained,  and  no  condensation  should  be 
allowed  to  accumulate.  High  and  low-pressure  drips  should  be  kept 
separate.  All  tanks  should  have  gauge  glasses. 

Fig.  331  shows  the  ap- 
plication, of  a  float  trap 
for  automatically  returning 
water  to  the  boiler.  For 
this  purpose  the  trap  must 
be  placed  three  feet  or 
more  above  the  water  line 
in  the  boiler,  so  that  the 
water  may  gravitate  to  it. 
Water  is  forced  into  the 
trap  from  the  returns 
through  pipe  A  until  it 
reaches  a  level  where  the 
float  opens  the  equalizing 
valve  V  and  permits  steam 
from  the  boiler  to  enter  the 
trap,  thus  equalizing  the 
pressures.  The  water  then 
flows  into  the  boiler  by 

gravity  through  check  valve  D.  At  the  end  of  discharge  the  float 
closes  the  equalizing  valve  and  another  cycle  begins.  Check  valve  C 
prevents  the  water  from  being  forced  back  to  the  return  pipe.  If  the 
pressure  in  the  return  pipe  A  is  not  sufficient  to  force  the  water  into 


FIG.  331.     Return  Trap. 


SEPARATORS,  TRAPS,  DRAINS 


597 


the  trap,  a  pump  or  another  trap  may  be  used  to  effect  this  result. 
Practically  any  high-pressure  trap  may  be  converted  into  a  return  trap 
by  proper  installation  and  an  "  equalizing  "  valve. 


FIG.  332.     Drainage  System  for  Jackets  and  Receivers  of  Triple  Expansion  Pumping 

Engines. 

Figs.  332  and  333  show  different  applications  of  steam  traps  to  the 
receiver  coils  and  jackets  of  triple  expansion  pumping  engines.  The 
drawings  are  self-explicit. 


FIG.  333.    Drainage  System  for  Jackets  and  Receivers  of  Triple  Expansion  Pumping 

Engines. 

319.  Drips  under  Vacuum.  —  Conditions  frequently  make  it  neces- 
sary to  remove  condensation  from  apparatus  working  under  a  vacuum, 
as,  for  example,  a  primary  heater. 

The  simplest  method  is  to  pipe  the  drips  to  the  condenser  and  per- 
mit the  condensation  to  gravitate  to  it  as  in  Fig.  334.  Where  this  is 
impracticable,  as  in  an  installation  with  the  condenser  above  the 


598 


STEAM   POWER  PLANT  ENGINEERING 


heater,  a  steam  trap  is  usually  employed.     Fig.  335  shows  the  applica- 
tion of  a  Bundy  trap  to  a  vacuum  or  primary  heater.     A  close-fitting 

weighted  check  valve  W,  set  to 
open  outwards  prevents  intake 
of  air  through  the  discharge 
pipe  while  the  trap  is  filling. 
Connection  E  is  made  from 
the  vent  underneath  the  valve 
stem  V  back  to  the  heater  so 
as  to  equalize  the  pressures. 
The  operation  is  as  follows: 
Condensation  gravitates  from 
the  heater  through  check  C  to 
'  the  body  of  the  trap,  the  check 

FIG.  334.     Gravity  Drainage;  Vacuum  Heater.     W    being    closed.         When     the 

bowl  is  full  enough  to  overcome 

the  weight  of  the  counterbalance,  it  sinks  and  opens  up  the  live-steam 
valve  V.    This  admits  steam  to  the  trap  through  pipe  D,  which  in 


FIG.  335.    Method  of  Draining  Heater  under  Vacuum. 

turn  closes  check  C  and  forces  the  water  past  the  weighted  check  W 
to  the  discharge  tank.  "  After  the  water  is  discharged  the  bowl  returns 


SEPARATORS,  TRAPS,  DRAINS 


599 


to  its  original  position  and  closes  valve  V,  the  weight  closes  check  W, 
the  vent  check  equalizes  the  pressure  in  the  bowl  and  heater,  and 
condensation  gravitates  to  the  trap  again. 

320.  Drips  under  Alternate  Pressure  and  Vacuum.  —  Occasionally 
the  load  on  an  engine  is  of  such  a  character  that  the  pressure  in  the 
receiver  alternates  from  a  pressure  of  30  or  40  pounds  absolute  to  a 
vacuum  of  varying  degree.  Where  the  periods  of  vacuum  operation 
are  very  few  and  of  short  duration,  as  in  the  average  installation,  no 
attention  is  paid  to  the  vacuum  and  the  condensation  is  removed  by 
a  trap  in  the  ordinary  way.  If,  however,  the  periods  are  of  sufficient 
duration  and  frequency,  the  ordinary  method  is  not  applicable  and  the 
arrangement  shown  in  Fig.  336  may  be  used.  The  trap  is  placed 


FIG.  336.     Method  of  Draining  Receivers  under  Alternate  Vacuum  and  Pressure. 

below  the  receiver  as  indicated.  The  delivery  pipe  is  provided  with  a 
weighted  check  or  resistance  valve  W  set  so  as  to  open  outwards  from 
the  trap,  also  a  spring  water  relief  valve  R.  Another  weighted  check  P 
is  placed  in  the  line  leading  from  the  vent  to  the  atmosphere,  and  a 
plain  check  C  in  the  line  leading  back  into  the  receiver.  This  arrange- 
ment of  valves  permits  the  venting  of  the  trap  after  discharge  and 
effectually  excludes  air  from  the  trap  when  there  is  less  than  atmos- 
pheric pressure  on  the  receiver.  With  the  relief  valve  set  to  open  at  a 


600 


STEAM  POWER  PLANT  ENGINEERING 


pressure  in  excess  of  the  maximum  receiver  pressure  it  acts  as  a 
"  stop  "  in  the  pipe  and  the  water  must  enter  the  trap.  When  the 
trap  discharges,  the  live  steam  supplied  through  the  pipe  attached  to 
the  steam  valve  forces  the  water  through  the  weighted  check  and  relief 
valves  into  the  sewer  or  receiving  tank.  When  working  with  a  vacuum, 
the  pressures  in  receiver  and  trap  are  equalized  through  the  vent  con- 
nection and  the  condensation  flows  into  the  trap  by  gravity.  The 
operation  of  discharge  is  the  same  as  in  the  case  of  pressure. 

321.   The  Steam  Loop.  —  Fig.  337  illustrates  the  principles  of  the 
"  steam  loop  "  for  automatically  returning  high-pressure  drips  to  the 


»X^_J 

FIG.  337.     General  Arrangement  of  the  Simple  "  Steam  Loop.*' 

boiler.  In  the  figure  the  loop  is  returning  the  condensation  from  a 
steam  separator  to  a  boiler  above  the  level  of  the  separator.  The 
apparatus  is  very  simple,  consisting  of  a  horizontal  and  two  vertical 
lengths  of  plain  pipe  placed  as  indicated.  Pipes  R  and  B  may  be  cov- 
ered but  "  horizontal"  A  is  left  uncovered,  as  its  function  is  that  of  a 
condenser.  The  operation  is  as  follows:  Circulation  is  first  started  by 
opening  stop  valve  0  at  the  bottom  of  the  drop  leg  until  steam 
escapes.  The  valve  is  then  closed  and  the  steam  in  the  horizontal  A 
condenses  and  gravitates  to  the  drop  leg  B.  On  account  of  the  slight 
reduction  in  pressure  in  the  horizontal  a  mixture  of  spray  and  steam 
flows  from  the  separator  chamber  to  the  horizontal,  and,  condensing, 
gravitates  to  the  drop  leg.  The  column  of  water  in  the  drop  leg  rises 
until  its  static  head  balances  the  difference  in  pressure  in  the  riser  R 
and  the  horizontal.  In  other  words,  a  decrease  in  pressure  in  the 
horizontal  produces  similar  effects  on  the  contents  of  the  riser  and 
drop  leg  but  in  a  degree  inversely  proportional  to  their  densities. 


SEPARATORS,  TRAPS,  DRAINS 


601 


602  STEAM  POWER  PLANT  ENGINEERING 

Any  further  accumulation  causes  an  equal  amount  to  pass  from  the 
bottom  of  the  column  to  the  boiler,  since  the  pressure  in  the  boiler  is 
then  less  than  that  at  the  bottom  of  the  column,  that  is,  the  steam 
pressure  on  the  top  of  the  water  column  plus  the  hydrostatic  head  H 
is  greater  than  the  pressure  in  the  boiler.  Once  started  the  process  is 
continuous  and  requires  no  further  attention. 

322.  The  Holly  Loop.  —  In  the  application  of  the  steam  loop  where 
many  points  requiring  drainage  are  connected  to  many  boilers  and 
conditions  are  more  complex,  some  method  other  than  the  simple  one 
of  radiation  may  be  advisable  to  secure  the  necessary  lower  pressure  at 
the  top  of  the  loop.     Such  a  method  is  illustrated  in  Fig.  338.     This 
arrangement  differs  from  the  simple  loop  in  that  all  condensation  first 
gravitates  to  a  "  Holly  "  receiver  (shown  in  detail  in  Fig.  339)  before 

passing    into   the    "  riser." 
j_  The  receiver  is  placed   be- 

low the  lowest  point  to  be 
drained  and  serves  as  a 
storage  for  large  or  unusual 
quantities  of  water  and  en- 
ables the  riser  to  act  at  a 
constant  rate  independent 
FIG.  339.  Holly  Receiver.  of  variable  discharge  into 

the  receiver.  Furthermore,  the  lower  pressure  in  the  discharge  cham- 
ber necessary  to  secure  the  lifting  of  the  mingled  steam  and  water 
through  the  riser,  instead  of  being  created  by  condensation  as  in  the 
simple  loop,  is  produced  by  a  reducing  valve  B  discharging  into  the 
feed-water  heater.  The  operation  of  the  Holly  loop  is  as  follows: 
Circulation  is  started  by  opening  valve  D  until  steam  appears.  Valve  D 
is  then  closed  and  the  reducing  valve  is  put  into  commission.  Condensa- 
tion from  separators,  traps,  and  pipes  gravitates  to  the  "  receiver,"  from 
which  it  is  forced  into  the  "  riser  "  in  the  form  of  a  spray.  The  spray- 
ing effect  is  produced  by  a  series  of  holes  drilled  in  pipe  A,  Fig.  339. 
From  this  receiver  the  spray  and  moisture  rise  to  the  "  discharge 
chamber,"  on  account  of  the  lower  pressure  at  that  point,  where 
the  steam  and  entrained  water  are  separated,  the  water  gravitat- 
ing to  the  bottom  of  the  chamber  and  thence  to  the  drop  leg,  and  the 
steam  discharging  through  the  reducing  valve  into  the  heater.  The 
principles  of  operation  are  exactly  the  same  as  in  the  simple  steam 
loop. 

323.  Returns  Tank  and  Pump.  —  Low-pressure  drips  in  connection 
with  heating  systems  may  be  returned  to  the  boiler  along  with  the 
condensation  from  the  heating  system  by  a  combined  pump  and  receiver 


SEPARATORS,  TRAPS,  DRAINS 


603 


as  shown  in  Fig.  340.  The  height  of  water  in  the  tank  controls  the 
operation  of  the  pump  through  the  medium  of  a  float  and  throttle  valve. 
This  combination  of  float  and  balanced  throttle  valve  is  sometimes 
called  a  "  pump  governor."  In  the  illustration  the  pump  forces  the 
returns  through  a  closed  heater  before  delivering  them  to  the  boiler, 


FIG.  340.    Returns  Tank  and  Pump. 

though  they  are  oftentimes  returned  directly.  The  tank  is  vented  to 
the  atmosphere  to  prevent  it  from  becoming  "  air  bound."  The  cold- 
water  supply  or  make-up  water  is  sometimes  discharged  into  the  receiv- 
ing tank  as  indicated.  With  open  heaters  the  cold  supply  is  ordinarily 
controlled  by  a  float  within  the  heater  itself. 

324.  Office  Building  Drains.  —  In  the  power  plants  of  tall  office  build- 
ings the  public  sewers  are  often  above  the  basement  level,  and  it  is 
necessary  to  remove  all  liquid  wastes  mechanically. 

The  Shone  pneumatic  ejector  has  been  found  to  serve  this  purpose 
effectually.  This  apparatus  is  placed  in  a  pit  in  the  basement  floor 
into  which  all  sewage,  drips  from  engines,  washings  from  boilers,  and 
ground  water  gravitate,  and  are  automatically  discharged  into  the 
street  sewer  by  means  of  compressed  air. 


604 


STEAM   POWER  PLANT  ENGINEERING 


Fig.  341  gives  a  sectional  view  of  a  Shone  ejector  of  ordinary  con- 
struction. It  consists  essentially  of  a  closed  vessel  furnished  with 
inlet  and  discharge  connections  fitted  with  check  valves,  A  and  B, 
opening  in  opposite  directions  with  regard  to  the  ejector.  Two  cast- 
iron  bells,  C  and  D,  are  linked  to  each  other,  in  reverse  positions,  the 


FIG.  341.    Shone  Ejector. 

rising  and  falling  of  which  control  the  supply  of  compressed  air  through 
the  agency  of  automatic  valve  E. 

The  bells  are  shown  in  their  lowest  position,  the  supply  of  compressed 
air  is  cut  off  from  the  ejector,  and  the  inside  of  the  vessel  is  open  to  the 
atmosphere.  The  sewage  gravitating  into  the  ejector  raises  the  bell  C, 
which  in  turn  actuates  the  automatic  valve  E,  thereby  closing  the  con- 
nection between  the  inside  of  the  ejector  and  the  atmosphere  and  open- 
ing the  connection  with  the  compressed  air.  The  air  pressure  expels 
the  contents  through  the  bell-mouthed  opening  at  the  bottom  and  the 
discharge  valve  B,  into  the  main  sewer.  Discharge  continues  until  the 
level  falls  to  such  a  point  that  the  weight  of  the  sewage  retained  in 
the  bell  D  is  sufficient  to  pull  it  down,  thereby  reversing  the  automatic 
valve.  This  cuts  off  the  supply  of  compressed  air  and  reduces  the 
pressure  to  that  of  the  atmosphere. 


SEPARATORS,  TRAPS,  DRAINS  605 

The  positions  of  the  bells  are  so  adjusted  that  compressed  air  is  not 
admitted  until  the  ejector  is  full,  and  is  not  allowed  to  exhaust  until 
emptied  down  to  the  discharge  level;  thus  the  ejector  discharges  a 
fixed  quantity  each  time  it  operates. 

Two  ejectors,  each  of  a  capacity  suitable  for  handling  the  average 
flow  of  tributary  sewage  and  so  arranged  that  they  can  work  either 
independently  or  together,  are  usually  installed  at  each  ejector  station. 

The  main  sanitary  sewer  of  the  building  usually  discharges  directly 
into  the  ejectors,  the  surface  water,  drips,  etc.,  being  collected  in  a 
neighboring  sump.  The  latter  is  connected  to  the  sanitary  sewer 
through  a  trap  or  back-water  valve. 


CHAPTER  XV. 

PIPING  AND  PIPE  FITTINGS. 

325.  General.  —  The   advent   of  high   pressures    and   superheat   is 
responsible  for  the  elimination  of  many  of  the  older  systems  of  piping, 
the  tendency  being  towards  greater  uniformity  in  design,  particularly 
in  electric  central  station  work.     In  isolated  stations  the  conditions 
of  operation  and  installation  are  so  variable  that  each  case  presents  an 
entirely  different  problem.     In  any  system  of  piping  the  fundamental 
object  is  to  conduct  the  fluid  in  the  safest  and  most  economical  manner. 

The  material  should  be  the  best  obtainable  and  the  system  so  flexi- 
ble that  a  break-down  in  one  element  will  not  necessitate  the  closing 
down  of  the  entire  plant.  On  the  other  hand,  flexibility  increases  the 
number  of  parts  and,  unless  first  cost  is  of  little  importance,  tends  to 
weaken  the  system  as  a  whole.  It  is  a  safe  general  proposition  to  say 
that  the  best  pipe  and  fittings,  irrespective  of  first  cost,  will  prove  the 
most  economical  in  the  end,  but  few  owners  of  power  plants  are  will- 
ing to  take  this  view. 

326.  Drawings.  —  An  assembly  drawing  of  the  entire  installation 
giving  the  location  of  all  valves  and  fittings  is  necessary  in  order  to 
avoid  interference,  and  particularly  where  a  number  of  fittings  are  to 
be  close  together.     Detailed  drawings  should  also  be  provided  of  each 
division  of  the  piping  to  facilitate  installation,  as,  for  example,  the 
high-pressure  steam,  the  exhaust  steam,  the  feed  water,  the  condensing 
water,  the  oil,  the  heating  and  the  sanitary  piping.     As  a  rule,  lower 
and  more  uniform  bids  will  be  obtained  from  an  isometric  or  perspec- 
tive sketch,  as  in  Fig.  254,  than  from  conventional  plan  and  elevation 
drawings,  due,  no  doubt,  to  the  greater  ease  with  which  the  drawing  is 
interpreted.    A  complete  set  of  specifications  for  a  piping  system  is  given 
in  paragraph  (415),  and  illustrates  the  usual  practice  along  this  line. 

327.  Materials  for  Pipes  and  Fittings.  —  The  following  materials  are 
used  in  the  construction  of  pipes  for  steam,  water,  and  gases. 

Average  Bursting  Tension. 

Low-carbon  or  mild  steel 65,000  Ibs.  per  sq.  in. 

Wrought  iron 50,000  Ibs.  per  sq.  in. 

Cast  iron,  high  grade 20,000  Ibs.  per  sq.  in. 

Cast  steel 50,000  Ibs.  per  sq.  in. 

Wrought  copper 33,000  Ibs.  per  sq.  in. 

Brass 18,000  Ibs.  per  sq.  in. 

Special  alloys  and  compounds 15,000-60.000  Ibs.  per  sq.  in. 

606 


PIPING  AND  PIPE  FITTINGS  607 

Mild  Steel.  —  The  greater  portion  of  the  piping  in  the  average  steam 
power  plant  is  of  mild  steel,  lap  or  butt  welded  for  high  pressures  and 
riveted  for  very  low  pressures  and  large  diameters.  Steel  pipe  is  con- 
siderably cheaper  than  that  manufactured  from  other  material,  and 
fulfills  practically  all  requirements  for  general  service. 

Wrought  Iron.  —  Pipes  manufactured  from  puddled  iron,  though  not 
as  strong  in  ultimate  bursting  strength  as  low-carbon  steel  pipes,  are 
superior  in  many  ways.  Threads  are  cut  more  readily  and  with  less 
power.  It  is  more  easily  bent  without  injury  and  offers  more  resistance 
to  corrosion.  The  life  of  wrought-iron  pipe  is  greater  than  that  of  steel 
pipe  under  conditions  of  extreme  exposure,  or  when  buried  under- 
ground, and  it  is  recommended  in  all  cases  where  corrosion  is  apt  to  be 
severe,  as  in  blow-off  pipes,  drips,  and  drains  or  in  pipes  not  in  continuous 
use.  Steam  pipes  well  covered  and  protected  from  external  moisture 
are  ordinarily  of  mild  steel,  as  the  conditions  do  not  warrant  the  use  of 
wrought  iron,  which  costs  approximately  10  per  cent  more.  Since  the 
term  "  wrought-iron  pipe  "  is  used  rather  loosely  in  practice,  the  manu- 
facturer will  generally  furnish  mild  steel  unless  special  stress  is  laid  upon 
the  term  "  wrought  iron." 

Wrought  Iron  vs.  Steel  Pipe:  IT.  Age,  March  2,  1905,  Jan.  18,  1906;  Pro.  Heat  and 
Vent.  Engr.,  Jan.,  1900;  Mach.,  Dec.,  1903,  p.  191;  Am.  Mfg.  and  Ir.  World,  April 
29,  1898;  Eng.  News,  50-292,  296,  487,  51-62;  Eng.  Rec.,  44-54;  Locomotive,  Jan. 
1,  1906. 

Corrosion  of  Pipes:   Trans.  A.S.M.E.,  18-282;  Eng.  Rec.,  42-194,  43-354,  45-584. 

Cast-Iron  Pipes.  —  Cast  iron  is  little  used  for  high-pressure  steam, 
except  occasionally  in  the  construction  of  headers  where  a  number  of 
branches  lead  into  a  single  pipe,  in  which  case  the  number  of  joints  is 
greatly  reduced  and  the  cost  considerably  less  than  for  wrought-iron  or 
steel  pipe  with  numerous  fittings  and  joints.  The  chief  objections  to 
cast  iron  for  high-pressure  steam  are  its  weight  and  lack  of  homogeneity. 
It  is  most  used  in  connection  with  water  service  and  sanitation. 

Cast-iron  Pipe:  Jour.  New  Eng.  Waterworks  Assn.,  March,  1907;  Engr.,  Lond., 
Nov.  7,  1902,  p.  454;  Power,  Jan.,  1904,  p.  334;  Eng.  News,  46-216,  48-193,  51-544. 

Cast-Steel  Pipe.  —  Cast-steel  headers  are  sometimes  used  in  power 
plants  for  highly  superheated  steam,  since  the  material  is  not  affected 
by  high  temperatures* to  the  same  extent  as  cast  iron.  High  first  cost 
and  the  difficulty  of  securing  castings  free  from  blow-holes  have  pre- 
vented its  more  general  use. 

Copper  Pipes.  —  Copper  steam  pipes  were  in  common  use  for  many 
years  in  marine  service  on  account  of  their  flexibility.  To  increase  the 
bursting  strength,  pipes  above  6  inches  in  diameter  were  generally  wound 


608  STEAM  POWER  PLANT  ENGINEERING 

with  a  close  spiral  of  copper  or  composition  wire.  In  recent  years 
wrought-iron  and  steel  pipe  bends  have  practically  superseded  copper 
for  flexible  connections.  As  a  rule  the  use  of  copper  pipes  should  be 
avoided  on  account  of  the  rapid  deterioration  of  the  metal  under  high 
temperatures  and  stress  variations.  The  cost  is  prohibitive  for  most 
purposes  and  this  alone  prevents  it  from  being  seriously  considered  in 
the  manufacture  of  pipe. 

Copper  Pipes:  Engr.,  Lond.,  April  15,  1898,  p.  360,  Aug.  11,  1893;  Engng,,  April, 
1898;  Eng.  Rec.,  July  30,  1898. 

Brass  Pipes.  —  Brass  is  little  used  in  the  construction  of  pipes  on 
account  of  its  high  cost.  It  withstands  corrosive  action  much  better 
than  iron  or  steel  and  is  often  used  in  connecting  the  feed  main  with 
the  boiler  drum.  Special  alloys,  nickel  steel,  "  ferrosteel,"  malleable 
iron,  and  the  like  have  been  used  in  the  manufacture  of  pipes,  and 
possess  points  of  superiority  over  wrought  iron  and  steel  for  some 
purposes,  as  for  highly  superheated  steam,  but  the  cost  is  prohibitive 
for  average  steam  power  plant  practice. 

Materials  for  Fittings.  —  Elbows,  tees,  flanges,  and  similar  fittings 
are  usually  made  of  cast  iron,  malleable  iron,  or  pressed  steel,  though 
cast  steel,  "  ferrosteel,"  and  other  steel  compounds  are  used  to  a  limited 
extent.  Standard  cast-iron  fittings  are  recommended  for  ordinary 
pressures  of  100  pounds  per  square  inch  or  less,  and  extra  heavy  cast- 
iron  fittings  for  higher  pressures.  Malleable  iron  fittings  are  lighter  and 
neater  than  cast  iron  and  are  extensively  used  for  small  sizes  of  steam 
and  gas  pipe. 

Manufacture  of  Pipe:  Sci.  Am.,  Dec.  12,  1903;  Mach.,  Feb.,  1904,  Dec.,  1903, 
p.  191;  Eng.  News,  50-232,  296. 

338.  Size  and  Strength  of  Commercial  Pipe.  —  Wrought-iron  and 
mild-steel  pipe  are  marketed  in  standard  sizes.  Those  most  commonly 
used  in  steam  power  plants  are  designated  as 

(1)  Merchant  or  standard  pipe. 

(2)  Full  weight  pipe. 

(3)  Large  O.  D.  pipe. 

(4)  Extra  heavy. 

(5)  Double  extra  heavy. 

Table  76  gives  the  dimensions  of  standard  "  full-weight  "  pipe, 
which  is  specified  by  the  nominal  inside  diameter  up  to  and  including 
12  inches  and  based  on  the  Briggs  standard.  Pipes  larger  than  12 
inches  are  designated  by  the  actual  outside  diameter  (O.  D.),  and  are 
made  in  various  weights  as  determined  by  the  thickness  of  metal  speci- 


PIPING  AND  PIPE  FITTINGS  609 

fied.  Manufacturers  specify  that  "  full  weight  "  pipe  may  have  a 
variation  of  5  per  cent  above  or  5  per  cent  below  the  nominal  or  table 
weights,  but  merchant  pipe,  which  is  the  standard  pipe  of  commerce, 
such  as  manufacturers  and  jobbers  usually  carry  in  stock,  is  almost 
invariably  under  the  nominal  weight.  It  varies  somewhat  among  the 
different  mills,  but  usually  lies  between  5  and  10  per  cent  under  the  table 
weight.  The  smaller  sizes  of  merchant  pipe,  J  inch  to  3  inches,  are  butt- 
welded  and  the  larger  sizes  are  lap-welded. 

Extra  heavy  and  double  extra  heavy  pipe  have  the  same  external 
diameter  as  the  standard,  but  are  of  greater  thickness  and  hence  the 
internal  diameter  is  smaller.  Taking  the  thickness  of  the  standard  pipe 
as  1,  that  of  the  extra  heavy  is  approximately  1.4  and  of  the  double 
extra  heavy  2.8. 

Wrought-iron  and  steel  pipes  are  ordinarily  designed  with  factors  of 
safety  of  from  6  to  15,  with  an  average  not  far  from  10.  The  standard 
hydrostatic  tests  to  which  the  various  pipes  are  subjected  at  the  mills 
are  as  follows: 

Hydrostatic  Pressure, 
Lbs.  per  Sq.  In. 

Standard,  butt-welded,  £-3  in 600  to  1,000 

Standard,  lap-welded,  3-12  in 500  to  1,000 

Extra  heavy,  butt-welded,  J-3  in 600  to  1,500 

Extra  heavy,  lap-welded,  1^-12  in 600  to  1,500 

Double  extra  heavy,  butt-welded,  |-2£  in 600  to  1,500 

Double  extra  heavy,  lap-welded,  1^-8  in 1,200  to  1,500 

The  pressure  necessary  to  burst  piping  is  far  above  anything  likely 
to  occur  in  ordinary  practice  on  account  of  the  thickness  of  material 
necessary  to  permit  of  threading,  thus: 

Actual  Bursting 

Pressure, 
Lbs.  per  Sq.  In. 

2-inch  "  standard  "  mild-steel  pipe* 5,800 

2-inch  wrought-iron  pipe 4, 106 

10-inch  "  mild-steel  pipef 3,000 

10-inch  wrought-iron  pipe  t 1,900 

10-inch  "  extra  heavy  "  wrought-iron  pipe  t 2,700 

*  Machinery,  December,  1903,  p.  192.  f  Crane  Company,  Published  Tests. 

Riveted  Pipes.  —  For  low  pressures  and  large  diameters,  pipes  are 
constructed  of  thin  sheets  of  boiler  steel  with  riveted  joints,  the  seams 
being  either  longitudinal  and  circumferential,  or  spiral.  Such  pipes 
are  not  necessarily  limited  to  large  sizes  and  low  pressures,  though  this 
is  the  usual  practice. 

Pipe  fittings  are  classed  as  screwed  or  flanged. 


610 


STEAM  POWER  PLANT  ENGINEERING 


329.  Screwed  Fittings,  Pipe  Threads.  —  For  screw  connections  the 
ends  of  pipes  and  fittings  are  threaded  to  conform  to  the  Briggs  or 
United  States  standard  system,  as  shown  in  Fig.  342.  The  end  of  the 


FIG.  342.    Standard  U.  S.  Pipe  Thread. 

pipe  is  tapered  1  to  32  with  the  axis,  the  angle  of  the  thread  being 
60  degrees  and  slightly  rounded  at  top  and  bottom.  The  proper  length 
of  perfect  threads  is  given  by  the  formula 


(146) 


in  which 


(0.8  D  +  4.8) 
n 

T  =  length  in  inches. 

D  —  actual  external  diameter  of  the  tube,  inches. 

n  =  number  of  threads  per  inch. 

The  imperfect  portion  of  the  thread  is  simply  incidental  to  the  pro- 
cess of  cutting.  The  object  of  the  taper  is  to  facilitate  "  taking  hold  " 
in  making  up  the  joint.  Table  76  gives  the  number  of  threads  per 
inch  for  various  sizes  of  standard  pipe.  When  properly  constructed  a 
screwed  joint  will  hold  against  any  pressure  consistent  with  the 
strength  of  the  pipe.  For  example,  the  ultimate  bursting  strength  of  a 
"  standard  "  2-inch  pipe  is  about  6000  pounds  per  square  inch,  while 
the  stripping  strength  of  the  joint  (with  perfect  threads)  is  225,000 
pounds.  The  threads,  however,  are  often  poorly  cut  and  the  parts 
screwed  together  improperly  cleaned  and  lubricated,  thus  causing 
leakage  between  the  threads. 

330.  Flanged  Fittings.  —  In  cast-iron  pipes,  valves,  tees,  and  other 
fittings  the  flange  is  always  a  part  of  the  casting,  but  for  joining 
the  two  ends  of  a  steel  or  wrought-iron  pipe  the  flanges  may  be 
fastened  to  the  pipe  in  a  number  of  ways.  Fig.  343,  A  to  H,  illus- 
trates methods  most  commonly  used.  In  A  to  C  the  pipes  are 
screwed  into  cast-iron  or  forged-steel  flanges  and  the  two  faces,  with 
metallic  or  composition  gasket  between,  are  drawn  together  by  bolts. 
A  illustrates  the  most  common  and  inexpensive  of  flanged  joints, 
which  requires  no  special  tools  and  can  be  made  up  at  the  place  of 


PIPING  AND  PIPE  FITTINGS 

•Maiog  jo  qoui  jad 


611 


I 


r-HOOlOt^l^CO 
IO  CO  t^  •*  (N  rH 

<N  --H 


ngth  of  Pipe  per 
Square  Foot  of 


II 


IOIOOO'—  l 


t^  Oi  CO  <M 
^H-^(X>Oi 
!>•  C^l  CO  "^  CO  O^  CO  O5  t^» 


s 


CO  i-H  t^  00  CO  CO 
t^rt<i-lTt<CO(NCOOO<r>T*<oOt>.          T-H 


rt< 
O 


iOOOSOCOCOOSCOiOOOOOOOCOCOOSOOCOrtl 


00  O5COOO 

COrtlCOCOCOOS 


03  O$  CO  CO  CO  Ti< 


OOO 


"S  «  "3  -S 

o  -S  g  <u 

P<    2    "g     Rj 

<       - 


-  « 

rt    C 


612 


STEAM  POWER  PLANT  ENGINEERING 


erection.  It  gives  satisfactory  results  for  pressures  of  100  pounds  or 
less,  but  for  higher  pressures  leakage  is  apt  to  take  place  between  the 
threads.  The  flanges  are  sometimes  made  with  a  long  thread  and  a 
recess  which  can  be  calked  with  soft  metal.  A  similar  joint  is  made 


RTTJ     A    km 


Rl  CT^ 


M  Eg 


SCREWED 


ROOVCD    FACCl 


SCREWED  &  PEENED 


SHRUNK  RIVETED 

FIG.  343.    Types  of  Pipe  Flanges. 


with  the  pipe  screwed  beyond  the  face  of  the  flange  and  the  two  faced 
off  together,  either  plane  or  as  shown  in  B,  which  is  known  as  a  male 
and  female  or  hydraulic  joint.  This  method  forms  a  very  reliable  joint, 
since  the  ends  of  the  pipe  bear  on  the  gasket,  and  the  gasket  is  pre- 


PIPING  AND  PIPE  FITTINGS  613 

vented  from  being  blown  out.  An  objection  lies  in  the  difficulty  of  open- 
ing the  line  to  remove  the  gasket  or  replace  a  fitting.  C  is  a  modifica- 
tion known  as  the  tongued  and  grooved  joint,  which  uses  an  extremely 
narrow  gasket.  Such  flanges  may  be  subjected  to  severe  strains  when 
the  bolts  are  drawn  up,  owing  to  the  small  area  of  contact.  Corru- 
gated copper  or  steel  gaskets  are  recommended,  since  soft  material  is 
apt  to  be  squeezed  out.  In  C  the  ends  of  the  pipe  are  peened,  which  is 
an  improvement  over  the  simple  screwed  joint.  D  illustrates  a  shrunk 
joint.  The  flanges  are  bored  for  a  shrink  fit  and  forced  over  the  pipe 
when  at  a  red  heat.  After  cooling,  the  end  is  beaded  over  into  a  recess 
on  the  face  of  the  flange  and  a  light  cut  taken  from  both.  H  shows  a 
modification  in  which  the  hub  is  riveted  to  the  pipe.  E  illustrates  a 
joint  constructed  by  rolling  the  pipe  into  a  corrugation  in  the  flange. 
The  end  of  the  pipe  is  then  faced  off  flush. 

One  of  the  best  commercial  joints  is  illustrated  by  F  and  is  known 
as  the  lap  joint.  The  pipe  is  expanded  as  indicated  and  a  light  cut  is 
then  taken  from  the  flared  ends  to  insure  a  tight  joint.  The  flanges 
are  loose  and  permit  of  considerable  flexibility  in  shifting  them  through 
various  angles.  This  is  sometimes  called  the  Van  Stone  joint. 

Pipes  with  flanges  welded  on  the  end  as  in  G  have  proved  the  most 
reliable  of  all  and  though  costly  are  considered  the  standard  for  high- 
pressure  and  high-temperature  work.  The  faces  are  ordinarily  raised 
sV  to  -jJg-  inch  inside  the  bolt  holes  and  ground  to  a  steam-tight  fit,  so 
that  thick  gaskets  are  unnecessary. 

For  moderately  high  pressures  and  temperatures  any  of  the  joints 
when  well  made  will  prove  satisfactory.  For  extremely  high  pres- 
sures and  temperatures  the  lap  or  welded  joints  are  preferable. 

The  comparative  costs  of  various  flanges  are  given  in  Table  77. 

On  July  18,  1894,  committees  of  the  American  Society  of  Mechan- 
ical Engineers,  of  the  National  Association  of  Steam  and  Hot  Water 
Fitters,  and  of  Manufacturers  met  and  adopted  a  schedule  for  dimen- 
sions of  flanges,  known  as  the  "  A.S.M.E.  Master  Steam  Fitters' " 
flange  schedule.  Flanges  dimensioned  in  accordance  with  this  schedule, 
Table  78,  are  applicable  for  pressures  up  to  100  pounds  per  square 
inch.  For  higher  pressures  a  schedule  shown  in  Table  79  was  adopted 
June  28,  1901. 

Pipe  Flanges  and  Fittings:  Power  &  Engr.,  Mar.  2,  1909,  p.  402,  June,  1904, 
p.  354,  Jan.,  1904,  p.  52,  Oct.,  1902,  p.  36;  Am.  Mach.,  April  18,  1901;  Jour.  Assn. 
Engr.  Soc.,  Sept.,  1904;  Jour.  Am.  Soc.  Nav.  Engrs.,  May,  1905;  Trans.  A.S.M.E., 
20-737,  429,  8-29,  347,  14-49,  7-311;  Eng.  News,  41-183,  323;  Eng.  Rec.,  41-113, 
44-209. 

Methods  of  Welding  Nozzles,  Flanges,  etc.,  to  Steam  Pipes:  Power  &  Engr.,  Sept.  28, 
1909,  p.  518. 


614 


STEAM  POWER  PLANT  ENGINEERING 


TABLE   77. 

COMPARATIVE  COST  OF   VARIOUS  PIPE  FLANGE  FITTINGS,    12-INCH  PIPE. 
(Circular  from  the  Crane  Company.) 


•  xJ 

rf^ 

-a 

C    3 

cj   x 

V 

"S 

M 

I  W 

'o  ^ 

0    = 

•a 

. 

m 

& 

a 

3 

£* 

"»  •£ 

• 

T3 

1 

£    > 

t£  '£> 

1 

02 

33 

^£ 

5s 

O 

£ 

s« 

02 

Cast  iron      

$7.40 

$16.00 

$18.00 

$13  00 

$21  00 

Ferrosteel      

8.70 

18.40 

20.00 

16.00 

23  40 

Malleable  iron  

9.90 

$22.00 

18.00 

Cast  steel 

22  40 

28  40 

34  00 

$33  00 

25  00 

33  40 

Weldless  steel 

26.40 

32.40 

38.00 

37.00 

$41.00 

30.00 

37.40 

Any  of  the  above  screwed,  shrunk,  welded,  rolled,  or  single-riveted  flanges  can  be 
furnished  with  male  or  female  face  at  $1 .25  extra. 

The  screwed  or  welded  flanges  can  be  furnished  with  tongued  or  grooved  face  at 
$1.25  extra. 

Any  of  the  above  screwed,  shrunk,  or  single-riveted  flanges  can  be  furnished  with 
calking  recess  at  $1.25  extra. 

TABLE   78. 

SCHEDULE  OF  STANDARD  FLANGES  ADOPTED  JULY  18,  1894,  BY  A  COM- 
MITTEE OF  THE  MASTER  STEAM  AND  HOT  WATER  FITTERS'  ASSOCIATION, 
AMERICAN  SOCIETY  OF  MECHANICAL  ENGINEERS,  AND  VALVE  AND 
FITTING  MANUFACTURERS.  SUITABLE  FOR  PRESSURES  UNDER  100 
POUNDS  PER  SQUARE  INCH. 


Size  of  Flange, 
Pipe 
SizeXDiam. 

Diameter 
of  Bolt 
Circle. 

Num- 
ber of 
Bolts. 

Size  of  Bolts, 
Pressure 
under  80 
Pounds. 

Size  of  Bolts, 
Pressure  80 
Pounds  and 
Over. 

Flange 
Thickness 
at  Hub  for 
Iron  Pipe. 

Flange 
Thickness 
at  Edge. 

Widl 
of 
Flani 
Faa 

2X6 

4| 

4 

|X2 

•-'^  1 

-X2 

1 

in. 

2 

2JX   7 

H 

4 

1x2^ 

i 

X21 

* 

21 

3  X   7* 

6 

4 

iX2ij 

1 

X2| 

21 

3*X  8* 

7 

4 

|x2i 

1 

X3J 

• 

i 

2* 

4X9 

7^ 

4 

iX2| 

\ 

X2i 

i 

2i 

4^X   91 

7| 

8 

^X3 

i 

X3 

. 

i 

2| 

5  X10 

4 

8 

X3 

i 

X3 

) 

2* 

6  Xll 

H 

8 

X3 

X3 

If 

1 

7  X12* 

10| 

8 

X31 

X31 

H 

1A 

2^ 

8  X13i 

8 

X  3^ 

IX  3i 

if 

ii 

2| 

9  X15 

131 

12 

f  X  3^ 

i 
. 

if 

it 

3 

10  X16 

141 

12 

1X3] 

|X3| 

2 

i^ 

3 

12  X19 

17 

12 

• 

2 

H 

3J 

14  X21 

18| 

12 

txii 

X4; 

2 

u 

15  X221 

20 

16 

1 

X41 

2 

u 

3J 

16  X23* 

211 

16 

1X4: 

1 

X41 

21 

i^s 

3i 

18  X25 

22| 

16 

1  X4^ 

11X41 



i^ 

3^ 

20  X27i 

25 

20 

1  X5 

HX5 



.  .  .  . 

5 

3J 

PIPING  AND  PIPE  FITTINGS 


615 


TABLE   79. 

SCHEDULE  OF  STANDARD  FLANGES  FOR  EXTRA  HEAVY  STEEL  PIPE,  FITTINGS, 
AND  VALVES  ADOPTED  JUNE  28,  1901,  BY  VALVE  AND  FITTING  MANUFAC- 
TURERS. SUITABLE  FOR  PRESSURES  FROM  125  TO  250  POUNDS  PER  SQUARE 
INCH. 


Size  of  Pipe, 
Inches. 

Diameter  of 
Flange, 
Inches. 

Thickness  of 
Flange, 
Inches. 

Diameter  of 
Bolt  Circle, 
Inches. 

Number  of 
Bolts. 

Diameter  of 
Bolts, 
Inches. 

2 

H 

* 

5 

4 

f 

21 

7* 

1 

5i 

4 

3 

It 

6^ 

8    • 

3* 

9 

7- 

8 

4 

10 

H6 

7- 

8 

4| 

10* 

1ft 

8 

8 

5 

11 

if 

9 

8 

6 

12* 

10 

12 

7 

14 

if 

Hi 

i 

12 

8 

15 

if 

13 

12 

9 

16 

if 

14 

12 

10 

in 

If 

15i 

16 

12 

20 

2 

171 

16 

14 

22* 

2| 

20 

20 

15 

23* 

21 

20 

1 

16 

25 

2f 

22i 

t 

20 

1 

18 

27 

2\- 

24* 

24 

1 

20 

29* 

21 

26| 

24 

1  J 

22 

31* 

2] 

28| 

28 

11 

24 

34 

2! 

3H 

28 

H 

TABLE   80. 

LOSS    OF    HEAT    FROM    BARE    STEAM    PIPE.* 
Still  Air. 


1      ® 

i 

*0    «H 

g 

| 

•8 

. 

o1  p,  <i 

Authority  of 

Descriptive  Refer- 

a 

|1 

1 

I 

i 

o| 

5l  S 

1^*1 

Test. 

ences. 

s 

•8 

CEJ    tic 

2T-S 

if 

S 

1  * 

S  -fa 

^  S 

ill 

.  fej  o 
&  a    . 

QQ 

02* 

5  O 

S  ** 

£  ^ 

5^ 

|S£ 

pjSS 

Barrus  
Do  
Do  

(Power,   Dec.,    1901;! 
iTrans.A.S.M.E.,  vol.1 
Ixxin;  Stevens  Ind.,| 
(Vol.  xix,  p.  388.       J 

2 
2 
10 

63.57 
63.92 
98.33 

82 
149 
149 

325 
365 
365 

56 
63.3 
73.6 

268.6 
302.2 
291.7 

0.915 
1.150 
1.085 

3.01 
3.25 
3.18 

Hudson  Beare 

3  53t 

8.13 

135 

358 

67 

291 

1.050 

3.10 

"1301bs."  

Stevens  Ind.,  Vol. 

U  .  VU  | 

2 

50.66 

128 

354 

80.1 

274.6 

0.994 

3.13 

xix,  p.  388. 

Jacobus  

Stevens  Ind.,  Vol. 

2 

7.63 

53 

301 

71.2 

229.6 

0.707 

2.78 

XVIII. 

Brill  

Trans   A  S  M  E 

g 

135.4 

110 

344 

75.5 

269 

0.834 

2.71 

Vol.'  XVL' 

*  C.  P.  Paulding,  Stevens  Indicator,  Vol.  xix,  p.  388.        t  Outside  diameter. 


616 


STEAM  POWER  PLANT  ENGINEERING 


331.  Coverings.  —  Steam  pipes,  feed-water  pipes,  boiler  steam 
drums,  receivers,  separators,  etc.,  should  be  covered  with  heat-insu- 
lating material  to  reduce  radiation  losses  to  a  minimum.  For  most 
practical  purposes  the  loss  of  heat  from  a  bare  steam  pipe  or  drum 
may  be  taken  as  3  B.T.U.  per  square  foot  per  hour  per  degree  differ- 
ence in  temperature,  Table  80.  The  actual  loss  depends  upon  the 
diameter  of  the  pipe,  on  its  position  whether  vertical  or  horizontal,  the 
nature  of  the  surface,  and  the  velocity  of  the  surrounding  air  currents. 
For  a  detailed  analysis  of  these  various  influences,  and  interesting 
information  on-  the  transmission  of  heat,  the  reader  is  referred  to 
Paulding's  "  Steam  in  Covered  and  Bare  Pipes." 

By  properly  applying  any  good  commercial  covering,  from  75  per 
cent  to  90  per  cent  of  the  heat  loss  may  be  prevented.  (See  Fig. 
344  and  Table  81.) 


FIG.  344.    Efficiency  of  Pipe  Coverings. 

Example:  Required  the  saving  per  annum  due  to  covering  a  pipe 
10  inches  in  diameter  and  100  feet  long;  steam  pressure  150  pounds; 
average  temperature  of  the  air  76  degrees  F. ;  cost  of  covering  applied 
65  cents  per  running  foot;  efficiency  of  covering  85  per  cent;  cost  of  coal 
$2.50  per  ton;  plant  to  operate  14  hours  per  day  and  300  days  per  year. 

The  temperature  of  steam  at  150  pounds  pressure  =  366  degrees  F. 


PIPING  AND  PIPE  FITTINGS 


617 


Difference  of  temperature  between  the  steam  and  air  =  366  —  76 
=  290  degrees  F. 

Loss  per  square  foot  per  hour,  bare  pipe  =  3  X  290  =  870  B.T.U. 

Loss  per  square  foot  per  day,  bare  pipe  =  870  X  14  =  12,180  B.T.U. 

Loss  per  square  foot  per  year,  bare  pipe  =  12,180  X  300  =  3,654,000 
B.T.U. 

100  lineal  feet  of  10-inch  pipe  has  an  external  surface  of  282  square 
feet.     Therefore  the  loss  per  year  from  the  bare  pipe  is 

282  X  3,654,000  =  1,030,000,000  B.T.U.  (approx.). 

TABLE  81. 

EXPERIMENTS    ON    STEAM-PIPE    COVERINGS. 
("Condensation  of  Steam  in  Covered  and  Bare  Pipes  "  [Paulding].) 


Kind  of  Covering. 

Diam. 
of  Test 
Pipe, 
Inches. 

Thick- 
ness of 
Cover- 
ing, 
Inches. 

Temperatures 
F. 

B.T.U.  per 
Hour  per 
Square  Foot 
of  Pipe 
Surface. 

Date 
of 
Test. 

Testing  Ex- 
pert. 

Steam. 

Air. 

Total. 

Per 
Degree 
Differ- 
ence. 

Hair  felt  

2 
8 
2 

2 

8 
8 
8 
2 
10 
2 
4 
2 
8 
2 
2 
10 
2 
2 
10 
8 
2 
4 
4 
2 
8 
2 
2 
8 
8 

0.96 
0.82 
0.88 

1.30 

1.30 
1.44 
.60 
.125 
.375 
.14 
.12 
.09 
.25 
.08 
.00 
.19 
.20 
.125 
.375 
.70 
.31 
.25 
.12 
.96 
.30 
.00 
0.99 
0.75 
0.75 

302.8 
348.3 
304.5 

306.6 

344.1 
346.1 
344.1 
364.8 
364.8 
309.2 
388.0 
354.7 
344.1 
310.9 
365.2 
365.2 
309.2 
365.2 
365.2 
345.5 
354.7 
388.0 
388.0 
303.3 
344.7 
354.7 
307.4 
347.1 
347.9 

71.4 
69.0 
73.3 

76.1 

58.3 
74.3 
63.0 
60.7 
62.8 
79.4 
72.0 
80.1 
66.3 
81.6 
64.6 
66.0 
79.4 
64.6 
66.8 
78.3 
80.1 
72.0 
72.0 
72.3 
79.0 
80.1 
72.5 
75.3 
74.3 

89.6 
117.9 
100.3 

83.7 

81.3 
86.1 
72.0 
145.0 
85.0 
59.7 
147.0 
155.8 
106.6 
69.8 
155.0 
103.0 
69.9 
176.0 
112.0 
93.4 
157.0 
143.0 
166.0 
165.5 
133.5 
198.0 
180.0 
238.0 
260.0 

0.387 
0.422 
0.434 

0.363 

0.284 
0.317 
0.256 
0.477 
0.248 
0.260 
0.465 
0.567 
0.384 
0.304 
0.515 
0.347 
0.304 
0.585 
0.375 
0.394 
0.572 
0.453 
0.525 
0.716 
0.502 
0.721 
0.766 
0.876 
0.950 

1901 
1894 
1901 

1901 

1894 
1894 
1894 
1901 
1901 
1901 
1896 
1896 
1895 
1901 
1901 
1901 
1901 
1901 
1901 
1894 
1896 
1896 
1896 
1901 
1894 
1896 
1901 
1894 
1894 

Jacobus 
Brill 
Jacobus 

Jacobus 

Brill 
Brill 
Brill 
Barrus 
Barrus 
Jacobus 
Norton 
Paulding 
Brill 
Jacobus 
Barrus 
Barrus 
Jacobus 
Barrus 
Barrus 
Brill 
Paulding 
Norton 
Norton 
Jacobus 
Brill 
Paulding 
Jacobus 
Brill 
Brill 

Do  
Remanit  for  interme- 
diate pressure. 
Remanit  for  high  pres- 
sure. 
Mineral  wool      

Champion  mineral  wool 
Rock  wool       

Asbestos  sponge  felted 
Do   

Do  

Magnesia   

Do  

Do  

Do  

Do  

Do  

Asbestos,  Navy  Brand 
Do  

Do  

Manville  sectional.  .  .  . 
Do 

Do  

Asbestos  air  cell  .  . 
Do.. 

Asbestos  fire  felt  
Do  

Do  

Fossil  meal.  

Riley  cement  

618  STEAM  POWER  PLANT  ENGINEERING 

Assuming  a  net  available  heat  value  of  10,000  B.T.U.  per  pound  for 
the  coal,  the  equivalent  coal  consumption  is  51.5  tons,  valued  at 
51.5  X  $2.50  =  $128.75. 

The  covering  will  save  85  per  cent  of  this,  or  $109.50  per  annum. 

The  pipe  covering  applied  will  cost  100  X  0.65  =  $65.00. 

In  this  case  the  covering  will  pay  for  itself  in  considerably  less  than 
a  year. 

Pipe  covering  is  applied  in  sections  molded  to  the  required  form 
and  held  to  the  pipe  by  bands,  or  may  be  applied  in  a  plastic  form. 
The  former  is  more  readily  applied  and  removed,  and  is  usually 
adopted  for  pipes,  while  the  valves  and  fittings  are  sometimes  covered 
with  plastic  material.  Piping  should  be  tested  under  pressure  before 
being  covered,  since  leaks  destroy  the  efficiency  and  life  of  the  cover- 
ing. If  the  surrounding  atmosphere  is  moist  the  covering  should  be 
given  two  or  three  coats  of  good  paint.  Coverings  are  sometimes 
applied  to  cold-water  pipe  to  prevent  sweating  in  a  humid  atmosphere. 

Pipe  Coverings:  Power,  July,  1904,  p.  407,  Aug.,  1904,  p.  482,  May,  1903,  p.  239, 
Dec.,  1901,  p.  32;  St.  Ry.  Jour.,  Nov.  29,  1902,  p.  875;  Engr.,  Lond.,  May  27,  1904, 
p.  547;  Eng.  Review,  Nov.,  1898,  p.  15;  Am.  Elecn.,  May,  1903;  Engng.,  Aug.  7,  1903; 
Mech.  Engr.,  Nov.  25,  1905;  Elec.  World  and  Engr.,  April  6,  1901;  Stevens  Ind., 
Oct.,  1902;  Trans.  A.S.M.E.,  16-827,  23-791. 

Identification  of  Power  House  Piping  by  Colors:  Power  &  Engr.,  Apr.  26, 1910,  p.  752. 

332.  Expansion.  —  One  of  the  most  difficult  problems  in  the  design 
of  a  piping  system  is  the  proper  provision  for  expansion  and  contrac- 
tion due  to  change  in  temperature.  If  a  pipe  is  immovably  fixed  at 
both  ends  and  under  no  strain  when  cold,  and  the  temperature  is 
increased,  as  by  the  admission  of  steam,  it  is  subjected  to  a  com- 
pression proportional  to  the  rise  in  temperature  (within  the  elastic 
limit).  For  example,  a  6-inch  standard  extra  heavy  wrought-iron 
pipe  200  feet  long  at  66  degrees  F.,  if  heated  to  366  degrees  F.  (the 
temperature  corresponding  to  steam  at  165  pounds  per  square  inch 
absolute  pressure),  will  exert  an  axial  force  of 

P  =  EA  d  -  t)fi.     (Mechanics  of  Engng.,  Church,  p.  218.)      (147) 

P  =  force  in  pounds. 

E  =  modulus  of  elasticity,  30,000,000. 

^  =  final  temperature,  degrees  F. 
t  =  initial  temperature. 

fi  =  coefficient  of  expansion,  0.0000075. 

A  =  sectional  area  of  the  pipe  material,  8.5  square  inches. 
Hence 

P  =  30,000,000  X  8.5  (366  -  66)  0.0000075. 
=  573,750  pounds. 


PIPING  AND  PIPE  FITTINGS 


619 


Unless  well  braced  throughout  its  entire  length  the  pipe  will  buckle 
and  become  distorted.  If  free  to  expand  its  length  would  increase. 
The  temperature  of  the  pipe  is  always  less  than  that  of  the  steam  on 
account  of  radiation  from  the  outer  surface  and  varies  with  the  effi- 
ciency of  the  covering.  But  ignoring  radiation  the  increase  in 
length  is 

l  =  p(tt-t)  L,  (148) 

in  which 

I  =  increase  in  length,  inches. 
L  =  length  of  pipe,  inches. 

Other  notations  as  in  (147). 

Substituting  in  (148),  tL  =  366. 
t   =  66. 
p  =  0.0000075. 
L  =  2400. 

I  =  0.0000075  (366  -  66)  2400 
=  5.4  inches. 

Since  the  forces  produced  by  expansion  are  practically  irresistible, 
the  pipe  is  invariably  allowed  to  expand  freely  by  suitable  means  so 
as  not  to  strain  the  connections.  The  coefficients  of  expansion  per 
degree  difference  in  temperature  for  various  pipe  materials  are  given 
in  Table  82. 

Headers  less  than  50  feet  in  length  usually  require  no  special  pro- 
visions for  expansion  provided  the  ends  are  free  and  the  leads  to  and 


FIG.  345.    Types  of  Expansion  Pipe  Bends. 

from  the  header  are  not  too  short,  the  pipe  usually  being  anchored  at 
the  middle  and  permitted  to  expand  in  either  direction.  Free  expan- 
sion of  the  feeders  may  be  provided  for 

(1)  By  long  radius  bends,  as  in  Fig.  345. 

(2)  By  double  swing  screwed  fittings,  as  in  Fig.  346,  or 

(3)  By  packed  expansion  joints,  Fig.  347. 

Where  practicable  the  long  radius  bends  will  prove  most  satis- 
factory. The  radius  of  the  bend  should  not  be  less  than  5  diameters 
of  the  pipe,  and  larger  if  possible.  The  length  of  straight  pipe  at  the 


620 


STEAM  POWER  PLANT  ENGINEERING 


end  of  each  bend  should  not  be  less  than  twice  the  diameter  of  the 
pipe  measured  from  the  face  of  the  flange. 

On  account  of  the  great  strains  to  which  the  joints  of  pipe  bends  are 
subjected,  the  welded  joint,  G,  Fig.  343,  is  recommended  as  giving  the 
best  results.  The  next  best  is  the  lap  joint,  F,  Fig.  343. 


TABLE  82. 

COEFFICIENTS    OF    LINEAR    EXPANSION    PIPING    MATERIALS. 


Material. 

Temperature 
Range. 

Mean  Coeffi- 
cient per  De- 
gree F. 

Wrought  iron  and  mild  steel 

32-212 

0  00000656 

Wrought  iron 

32-572 

0  00000895 

Cast  iron  

32-212 

0.00000618 

Cast  steel      

32-212 

0  00000600 

Hardened  steel  ,  

32-212 

0.00000689 

Nickel-steel,  36  per  cent  Nickel  

32-572 

0.00000030 

Copper,  cast  

32-212 

0.00000955 

Copper  wrought 

32-572 

0  00001092 

Lead 

32-212 

0  00001580 

Cast  brass 

32-212 

0  00001043 

Brass  wire  and  sheets           

32-212 

0.00001075 

Tin  cast                         

32-212 

0  00001207 

Tin  hammered     

32-212 

0.00001500 

Zinc  cast       

32-212 

0.00001633 

Zinc  hammered 

32-212 

0  00001722 

LINEAR    EXPANSION    OR    CONTRACTION    OF    CAST    IRON    IN    INCHES    PER 
100    FEET,  —  DEGREES    F. 


Temperature  Difference. 

Expansion. 

Temperature  Difference. 

Expansion. 

100 
150 
200 
250 

0.72 
1.1016 
1.5024 
1.9260 

300 
400 
500 
600 
800 

2.376 
3.360 
4.440 
5.616 

7.872 

Multiply  by  1.1  for  wrought  mild  steel. 
Multiply  by  1.5  for  wrought  copper. 
,     '  Multiply  by  1.6  for  wrought  brass. 

Fig.  345,  A,  B,  C,  D  shows  applications  of  pipe  bends  to  straight  pipe 
runs.  A  is  the  cheapest  and  most  common  arrangement  for  all  sizes  of 
pipe.  B  is  a  modification  for  limited  center  to  center  spaces.  C  shows 
a  common  method  of  taking  up  expansion  in  straight  runs  of  pipe  of 
very  large  diameters  where  the  space  requirements  prohibit  the  use  of 


PIPING  AND  PIPE  FITTINGS 


621 


a  single  U  bend.  Here  the  main  runs  are  connected  to  manifolds  which 
in  turn  are  connected  by  a  number  of  small  U  bends,  the  equivalent 
areas  of  which  correspond  to  that  of  the  large  pipes.  This  makes  a  more 
flexible  connection  than  if  a  single  U  bend  were  used.  The  arrange- 
ment D  does  away  with  the  elbows  required  in  A, 
but  is  not  applicable  to  pipes  over  8  inches  in 
diameter. 

Figs.  357  and  358  show  applications  of  pipe 

bends  to  boiler  and  header  connections. 

^         Ck^  ^>'  *^  shows  a  double  swing  screwed  joint 

•^>  in  which  expansion  causes  the  fittings   to   turn 

slightly  and 
thus  relieve  the 
strain.  This 
method  is  usu- 
ally adopted 
where  long  ra- 
dius bends  are 


FIG.  347.    Slip  Expansion  Joint. 


FIG.  346.  "Double-Swing"  not  practicable 

Expansion  Joint.  on    accoUnt     of 

lack  of  space  and  where  screwed  fittings  are  used. 

Slip  joints,  Fig.  347,  are  now  little  used  except  with  very  large  pipes 
and  where  space  prohibits  long  radius  bends.  When  slip  joints  are 
employed  the  pipe  must  be  securely  anchored  to  prevent  the  steam 
pressure  from  forcing  the  joint  apart  and  at  the  same  time  permit  the 
pipe  in  expanding  to  work  freely  in  the  stuffing  box.  Sagging  of  the 
pipe  on  either  side,  which  might  cause  binding  in  the  joint,  is  prevented 
by  suitable  supports. 

Expansion  in  Steam  Pipes:  Power,  July,  1906,  p.  426,  Jan.,  1904,  p.  30,  March, 
1904,  p.  160,  Oct.,  1904,  p.  609,  Dec.,  1900;  Am.  Elecn.,  10-432;  Engr.,  U.S., 
Feb.  1,  1904,  p.  125;  Eng.  News,  44-194,  47-468,  50-487;  Power,  June  2,  1908. 

333.  Pipe  Supports  and  Anchors.  —  Pipe  lines  must  be  supported  to 
guard  against  excessive  deflection  and  vibration.  Supports  are  con- 
veniently classified  as  (1)  hangers,  (2)  wall  brackets,  and  (3)  floor  stands. 

Fig.  348  illustrates  a  type  of  hanger  for  suspending  pipes  from  I 
beams.  The  supports  being  free  to  swing, "no  provision  for  expansion 
is  necessary.  A  properly  designed  hanger  may  be  readily  removed 
without  disturbing  the  pipe  line,  and  should  be  adjustable  to  facilitate 
"  lining  up."  If  of  rigid  construction  the  lower  end  should  be  provided 
with  a  roller. 

Fig.  349  gives  the  details  of  a  wall  bracket  with  rolls  and  roll  binder. 


622 


STEAM  POWER  PLANT  ENGINEERING 


Supports  adjacent  to  long  radius  bends  should  be  provided  with  roll 
binders  as  illustrated  to  prevent  the  pipe  from  springing  laterally,  but 


FIG.  348.    A  Typ- 
ical Pipe  Hanger. 


FIG.    349.      A    Typical     Wall 
Bracket  with  Binding  Roll. 


FIG.  349a.  A  Typ- 
ical Floor  Stand. 


they  may  otherwise  be  omitted.  The  rollers  are  often  made  adjustable 

to  facilitate  lining  up. 

Fig.  349a  illustrates  a  typical  floor  stand. 
Pipe  lines  are  usually  securely  anchored 
at  suitable  points  in  a  manner  similar  to 
that  illustrated  in  Fig.  350,  the  pipe  rest- 
ing on  a  saddle  and  being  rigidly  clamped 
to  the  bracket  by  a  flat  iron  band  with 
ends  threaded  and  bolted.  This  limits 
expansion  to  one  direction  and  prevents 
excessive  strain  on  the  fittings. 

334.  General  Arrangement  of  High- 
Pressure  Steam  Piping.  —  The  general  ar- 
rangement of  piping  depends  in  a  great 
measure  upon  the  space  available  for  en- 
gines and  boilers. 

The  engine  and  boiler  room  may  be 
placed 


FIG  360.  A  Typical  Pipe  Anchor. 


(1)   Back  to  back,  Fig.  361,  (2)    End  to  end,  Fig.  351, 

(3)    Double  decked,  Fig.  356. 


PIPING  AND  PIPE  FITTINGS 


623 


The  back  to  back  arrangement  is  the  most  common  and,  .other  things 
permitting,  is  to  be  preferred  on 
account  of  the  short  and  direct 
connection  between  engines  and 
boilers  and  the  ease  of  enlarge- 
ment. The  engine  and  boiler 
rooms  are  separated  by  a  wall, 
and  as  much  of  the  piping  as 
possible  is  located  in  the  boiler 
room. 

The  end  to  end  arrangement 
is  ordinarily  limited  to  situa- 
tions where  the  distribution  of 
space  precludes  the  back  to  back 
system. 

The  double  decked  arrangement 
is  frequently  used  where  ground 
space  is  limited  or  expensive. 

Engines  and  boilers  are  con- 
nected in  a  variety  of  ways 
through  steam  headers  as  shown 
in  the  following  examples: 

(1)  Single  header,  Fig.  361, 

(2)  Duplicate  headers,  Fig. 
352, 

(3)  Loop  or  ring  header,  Fig. 
353, 

(4)  The  "  unit "  system,  Fig. 
357. 

The  single  header  system  is  per- 
haps the  most  common,  since  it 
embodies  simplicity,  low  first 
cost,  and  provision  for  extension. 

The  duplicate  system  is  losing 
favor,  since  experience  shows 
that  the  extra  cost  of  the  du- 
plicate mains  will  usually  give 
better  returns  in  continuity  of 
operation  and  maintenance  if  in- 
vested in  high-grade  fittings  on 
a  single  pipe  system. 

The  loop  header  is  well  adapted  for  the  power  plants  of  tall  office  build- 


STEAM  POWER  PLANT  ENGINEERING 


3LL 


FIG.  352.    Typical  Duplicate  Header  System. 


FIG.  353.    Typical  Loop  Header. 


PIPING  AND  PIPE  FITTINGS 


625 


ings,  Fig.  355,  in  which  a  large  number  of  steam  engines,  elevator  pumps, 
air  compressors,  and  miscellaneous  steam-consuming  appliances  are 
crowded  together  in  a  comparatively  small  space. 

Large  modern  power  plants  are,  by  the  latest  practice,  divided  into 
complete  and  independent  units,  as  in  Fig.  440,  each  prime  mover 
having  its  own  boiler  equipment,  coal  and  ash-handling  machinery, 
feed  pumps,  and  piping,  operated  independently  of  the  rest  of  the  plant, 
though  provision  is  made  whereby  any  boiler  equipment  may  provide 
steam  for  any  prime  mover. 


FIG.  354.    Typical  By-Pass  System. 

The  power  plant  of  the  Manhattan  Elevated  Railway  Company, 
New  York,  is  practically  divided  into  eight  sections  each  consisting 
of  an  engine  and  eight  boilers,  the  boilers  being  "  double  decked " 
(Fig.  356). 

The  branch  pipes  from  the  upper  and  lower  batteries  lead  into  18- 
inch  headers,  the  steam  from  each  being  conducted  to  a  receiver  reservoir 
36  inches  in  diameter  and  20  feet  long  in  the  engine  room  basement 
directly  behind  each  engine,  from  which  the  two  high-pressure  cylinders 
are  supplied.  Gate  valves  are  used  in  each  boiler  branch,  one  close  to 
the  boiler  and  another  near  the  header,  and  also  in  the  steam  pipes 
near  the  reservoir.  The  steam  headers  for  each  of  the  eight  units  are 
connected  by  a  main  which  equalizes  the  pressure  and  allows  a  deficiency 
in  one  unit  to  be  made  up  from  the  others. 

Figs.  357  and  358  show  the  general  arrangement  of  the  steam  piping 
at  the  Yonkers  power  house  of  the  New  York  Central.  The  turbines 


626 


STEAM  POWER  PLANT  ENGINEERING 


PIPING  AND  PIPE  FITTINGS 


627 


628 


STEAM  POWER  PLANT  ENGINEERING 


Detail  Plan 


(Power) 


FIG.  358.    Details  of  Boiler  Steam  Piping,  Yonkers  Power  House  of  the    New  York 

Central  R.R. 


PIPING  AND  PIPE  FITTINGS 


629 


are  connected  in  pairs  by  14-inch  loops,  each  turbine  taking  steam 
from  either  of  two  banks  of  four  boilers.  The  high-pressure  steam 
piping  is  of  mild  steel  with  modified  reenforced  "  Van  Stone  "  joints. 
The  high-pressure  valves  are  of  the  split  disk  pattern  with  semi-steel 
bodies.  Expansion  is  taken  up  by  the  long  sweep  bends. 


FIG.  359.     Overhead  Boiler  Piping,  Quincy  Point  Power  Plant  of  the  Old  Colony  St. 
Ry.  Co.,   Quincy  Point,  Mass. 

Plants  using  superheated  steam  are  ordinarily  piped  to  supply 
saturated  steam  to  the  auxiliaries  as  illustrated  in  Fig.  359.  The 
boiler  branch  E,  leading  to  the  main  header,  normally  supplies  super- 
heated steam  to  the  engines.  C  is  an  auxiliary  main  supplying  the 
air  pumps,  stoker  engines,  and  other  auxiliaries  with  saturated  steam 
from  branch  pipe  D. 

Steam  Piping:  Power,  Feb.  23, 1909,  Nov.,  1905,  p.  683,  April,  1904,  p.  213,  Feb., 
1904,  p.  90,  Sept.,  1904,  p.  540,  Nov.,  1904,  p.  677,  July,  1903,  p.  356;  Ir.  Tr.  Rev., 
May  29,  1902;  Ir.  Age,  Dec.  11,  1902;  Met.  Work,  Feb.  14,  1903;  St.  Ry.  Rev.,  Nov. 
20,  1904,  p.  869;  St.  Ry.  Jour.,  Oct.  15,  1904,  p.  441 ;  Am.  Elecn.,  March,  1905,  p.  127; 
Elec.  Engr.,  Lond.,  Dec.  23,  1904;  Engr.  U.S.,  Feb.  15,  1905,  Dec.  1,  1904;  Elecn., 
Lond.,  July  21,  1899;  Eng.  Rec.,  48-90;  Trans.  A.S.M.E.,  15-536;  Cass.  Mag.,  June, 
1906. 

335.  Main  Steam  Headers.  —  Until  quite  recently  it  was  the  usual 
practice  to  make  the  area  of  the  steam  header  equivalent  to  the  com- 
bined areas  of  the  feeders,  but  the  function  of  the  header  is  now 
regarded  as  that  of  an  equalizer  rather  than  a  storage  reservoir.  In 
the  various  large  power  houses  recently  built  in  New  York  City,  with 
ultimate  capacities  of  from  60,000  to  150,000  kilowatts,  the  largest 
steam  headers  are  not  over  16  inches  in  diameter.  In  some  recent 


630 


STEAM   POWER  PLANT  ENGINEERING 


designs  the  pipes  leading  from  the  header  to  the  engines  are  two  sizes 
smaller  than  called  for  by  the  engine  builders.  In  this  case  large 
receiver  separators  two  to  four  times  the  volume  of  the  high-pressure 
cylinder  are  provided  near  the  throttle  as  in  Fig.  356.  The  pipes 
between  receiver  and  engine  are  full  size.  The  object  of  the  arrange- 
ment is  to  give  (1)  a  constant  flow  of  steam,  (2)  a  full  supply  of  steam 
close  to  the  throttle,  and  (3)  a  cushion  near  the  engine  for  absorbing 
the  shock  caused  by  cut-off.  With  moderately  superheated  steam  and 
boiler  pressures  from  125  to  150  pounds  a  velocity  of  8000  feet  per 
minute  is  allowed  in  the  header  and  as  high  as  9000  feet  per  minute 
between  header  and  receiver.  With  steam  turbines  velocities  as  high 
as  12,000  feet  per  minute  are  permissible,  provided  the  pipe  is  less  than 
50  feet  in  length  and  practically  free  from  sharp  bends.  Main  headers 
are  ordinarily  constructed  of  mild  steel,  though  cast-iron  and  cast- 
steel  headers  are  not  uncommon.  Cast  headers  permit  of  fewer  joints 
and  are  well  adapted  to  situations  where  a  number  of  branches  are 
closely  grouped  as  in  Fig.  361.  Cast-iron  headers  are  employed  in  the 
Manhattan  Elevated  Railway  power  station,  New  York. 


FIG.  360.     Steam  Header  and  Branches,  Grand   Rapids,   Grand   Haven  and  Muskegon 

Ry.  Co.  Power  House. 

The  proper  arrangement  and  number  of  valves  in  the  main  header 
and  feeders  has  been  a  subject  of  much  consideration.  Figs.  356  to 
361  show  some  of  the  different  successful  arrangements  in  recent 
installations.  In  Fig.  360  there  is  but  one  valve  between  the  boiler 
nozzle  and  the  main  header,  while  in  Fig.  361  there  are  two.  The 
latter  is  the  more  common  arrangement.  Where  two  valves  are 


PIPING  AND  PIPE  FITTINGS 


631 


632  STEAM  POWER   PLANT  ENGINEERING 

placed  in  a  feeder  they  should  be  arranged  so  as  not  to  form  a  pocket 
for  the  accumulation  of  leakage.  In  a  number  of  recent  installa- 
tions, Fig.  358,  the  valve  nearest  the  boiler  is  of  the  "  automatic  stop 
and  check  "  type,  its  function  being  the  automatic  cutting  off  of  the 
steam  from  the  header  should  the  pressure  in  the  boiler  suddenly  drop 
as  in  case  of  blowing  out  a  tube. 

Arrangement  of  Steam  Piping:  Am.  Elecn.,  April,  1905,  June,  1902,  p.  257,  June, 
1900;  Engr.  U.S.,  Dec.  1,  1904;  Mech.  Engr.,  Nov.  4,  1905;  Power,  Sept.,  1904,  p.  511, 
July,  1902;  Eng.  News,  Nov.  26,  1903,  p.  487;  Elec.  Rev.,  Lond.,  Aug.  11,  1899, 
p.  251;  St.  Ry.  Rev.,  Jan.,  1900,  p.  12;  Nov.,  1904,  p.  869. 

336.  Flow  of  Steam  in  Pipes.*  —  The  several  accepted  formulas 
relating  to  the  flow  of  steam  in  pipes  have  been  based  upon  a  few 
experiments  limited  to  pipes  of  small  diameter,  hence  the  application 
of  these  formulas  to  larger  pipes  or  to  conditions  other  than  those 
under  which  they  were  deduced  is  apt  to  lead  to  considerable  error. 
In  small  plants  extreme  accuracy  in  determining  the  proper  sizes  is 
not  necessary;  it  is  better  to  err  in  the  installation  of  too  large  a  pipe 
than  one  too  small.  In  larger  stations,  however,  where  the  pipes  are 
large  and  the  pressure  high,  the  cost  of  the  piping  increases  very 
rapidly  with  the  size.  For  example,  the  cost  of  10-inch  high-pressure 
fittings  is  from  15  to  20  per  cent  greater  than  9-inch  fittings,  and  in 
large  installations  this  first  cost  item  may  be  of  considerable  impor- 
tance. 

The  simplest  and  most  commonly  used  formula  is  based  upon  an 
allowable  steam  velocity  of  6000  feet  per  minute,  friction  and  other 
causes  of  drop  in  pressure  being  disregarded;  thus,  for  a  velocity  of 
6000  feet  per  minute, 

d  =  0.175        V,  (149) 


in  which 

d  =  diameter  of  the  pipe  in  inches, 

y=  density  of  the  steam  in  pounds  per  cubic  feet,  and 

W  =  weight  of  steam  flowing  in  pounds  per  minute. 

In  determining  the  diameter  of  the  steam  pipe  opening  for  recipro- 
cating engines  a  much  lower  velocity  than  6000  feet  per  minute  is 
assumed,  to  allow  for  the  various  conditions  of  operation.  Average 
practice  gives  the  constant  in  equation  (149)  a  value  of  0.3  instead  of 
0.175  when  used  in  this  connection. 

Equation  (149)  gives  satisfactory  results  for  pipes  under  100  feet  in 
length  and  between  4  and  8  inches  in  diameter;  for  larger  diameters 
the  velocity  could  be  increased  with  advantage;  for  smaller  diameters 

*  See  author's  original  paper,  Power,  June,  1907,  p.  377. 


PIPING  AND  PIPE  FITTINGS 


633 


or  greater  lengths  friction  and  condensation  would  cause  considerable 
drop  in  pressure  and  some  one  of  the  approved  formulas  in  Table  83 
should  be  used  instead. 


1000  2000   3000  4000   5000 


6000       SCCO       10000       12000 

Mean  Velocity,  Feet  per  Minute 


13000 


FIG.  362.    Drop  in  Pressure  for  Various  Velocities  and  Pipe  Sizes.     Initial  Pressure  100 
Pounds  Gauge,  Length  of  Pipe  100  feet. 

A  large  drop  in  pressure  means  a  small  pipe  and  high  velocity  with 
consequent  decrease  in  condensation,  but  a  point  is  soon  reached 
where  the  economy  in  the  size  of  pipe  is  more  than  offset  by  the  loss 
in  friction.  There  seems  to  be  no  fixed  rule  for  determining  the  drop 


634 


STEAM  POWER  PLANT  ENGINEERING 


Equivalent 
Units.  * 

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PIPING  AND  PIPE  FITTINGS 


635 


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636  STEAM  POWER  PLANT  ENGINEERING 

most  suitable  for  any  given  set  of  conditions.  In  current  practice  the 
drop  in  pressure  between  boiler  and  engine  ranges  from  a  fraction  of 
one  pound  to  four  pounds  per  square  inch  per  100  feet  of  pipe,  with  an 
average  between  one  and  two  pounds. 

Table  83  gives  a  few  of  the  best  known  formulas  for  the  flow  of 
steam,  and  Table  84  a  comparison  between  them  with  respect  to 
velocity,  weight  discharged,  diameter,  and  the  drop  in  pressure. 

Formula  11,  Table  83,  is  the  most  commonly  accepted,  and  the 
curves  in  Fig.  362  are  based  upon  it,  assuming  a  steam  pressure  of  100 
pounds  absolute  and  pipe  lengths  of  100  feet.  Within  the  limit  of 
12,000  feet  per  minute  velocity  and  10  pounds  per  square  inch  drop  in 
pressure  the  curves  are  sufficiently  accurate  for  all  practical  purposes, 
but  beyond  this  range  the  results  are  purely  conjectural  and  may  not 
be  accurate,  as  no  recorded  experiments  have  been  conducted  at  these 
high  velocities  or  with  pipes  of  large  diameters. 

Though  applicable  directly  to  pipes  100  feet  long  with  mean  pres- 
sure of  100  pounds  per  square  inch  absolute,  they  may  be  used  for  any 
length  or  pressure.  For  example,  for  any  length  other  than  100  feet, 
multiply  the  drop  given  in  the  curves  by  the  required  length  in  feet 
and  divide  by  100.  For  any  pressure  other  than  100  pounds  abso- 
lute, multiply  the  drop  given  in  the  curves  by  0.2271  (density  of  steam 
in  pounds)  and  divide  by  the  density  of  steam  at  the  required 
pressure. 

Table  85  is  the  table  ordinarily  used  in  connection  with  the  flow  of 
steam  and  is  calculated  from  equation  11.  Table  86  is  based  upon 
equations  4  to  12.  The  results  differ  slightly  from  those  in  Table  85, 
though  the  latter  is  more  comprehensive.  The  left-hand  half  of  Table 
86  gives  the  discharge  in  pounds  per  minute  for  pipes  of  various 
diameters  corresponding  to  drop  of  pressure  as  given  on  the  right-hand 
side  in  the  same  horizontal  line;  e.g.,  a  6-inch  pipe  100  feet  long  dis- 
charges 371  pounds  of  steam  per  minute  for  a  drop  of  16.4  pounds  at 
100  pounds  pressure. 

337.  Equation  of  Pipes.  —  It  is  frequently  desirable  to  know  what 
number  of  one  sized  pipes  will  be  equal  in  capacity  to  another  pipe. 

According  to  the  formulas  in  Group  II,  Table  84,  the  weights  dis- 
charged vary  with  the  square  root  of  the  fifth  power  of  the  diameter; 
that  is,  the  number  of  pipes  equal  in  capacity  to  any  given  pipe  may 
be  determined  from  the  equation 

N^dv+df,  (150) 

in  which  Nl  =  number  of  pipes  of  diameter  d1  equal  in  capacity  to  a 
pipe  of  diameter  d;  dt  and  d  in  inches. 


PIPING  AND  PIPE  FITTINGS 


637 


TABLE   85. 
FLOW    OF    STEAM    THROUGH    PIPES    (BABCOCK). 


Diameter  of  Pipe,  in  Inches.    Length  of  each  =  240  diameters. 

Initial  Pres- 

sure by 

Gauge. 

* 

1 

1* 

2 

2* 

3 

4 

Pounds  per 

Square  Inch. 

Weight  of  Steam  per  Minute,  in  pounds,  with  One  Pound  Loss  of  Pressure. 

1 

1.16 

2.07 

5.7 

10.27 

15.45 

25.38 

46.85 

10 

1.44 

2.57 

7.1 

12.72 

19.15 

31.45 

58.05 

20 

1.70 

3.02 

8.3 

14.94 

22.49 

36.94 

68.20 

30 

1.91 

3.40 

9.4 

16.84 

25.35 

41.63 

76.84 

40 

2.10 

3.74 

10.3 

18.51 

27.87 

45.77 

84.49 

50 

2.27 

4.04 

11.2 

20.01 

30.13 

49.48 

91.34 

60 

2.43 

4.32 

11.9 

21.38 

32.19 

52.87 

97.60 

70 

2.57 

4.58 

12.6 

22.65 

34.10 

56.00 

103.37 

80 

2.71 

4.82 

13.3 

23.82 

35.87 

58.91 

108.74 

90 

2.83 

5.04 

13.9 

24.92 

37.52 

61.62 

113.74 

100 

2.95 

5.25 

14.5 

25.96 

39.07 

64.18 

118.47 

120 

3.16 

5.63 

15.5 

27.85 

41.93 

68.87 

127.12 

150 

3.45 

6.14 

17.0 

30.37 

45.72 

75.09 

138.61 

Diameter  of  Pipe,  in  Inches.     Length  of  each  =240  diameters. 

Initial  Pres- 
sure by  Gauge. 

5 

6 

8 

10 

12 

15 

18 

Pounds  per 

Square  Inch. 

Weight  of  Steam  per  Minute,  in  Pounds,  with  One  Pound  Loss  of  Pressure. 

1 

77.3 

115.9 

211.4 

341.1 

502.4 

804 

1177 

10 

95.8 

143.6 

262.0 

422.7 

622.5 

996 

1458 

20 

112.6 

168.7 

307.8 

496.5 

731.3 

1170 

1713 

30 

126.9 

190.1 

346.8 

559.5 

824.1 

1318 

1930 

40 

139.5 

209.0 

381.3 

615.3 

906.0 

1450 

2122 

50 

150.8 

226.0 

412.2 

665.0 

979.5 

1567 

2294 

60 

161.1 

241.5 

440.5 

710.6 

1046.7 

1675 

2451 

70 

170.7 

255.8 

466.5 

752.7 

1108.5 

1774 

2596 

80 

179.5 

269.0 

490.7 

791.7 

1166.1 

1866 

2731 

90 

187.8 

281.4 

513.3 

828.1 

1219.8 

1951 

2856 

100 

195.6 

293.1 

534.6 

862.6 

1270.1 

2032 

2975 

120 

209.9 

314.5 

573.7 

925.6 

1363.3 

2181 

3193 

150 

228.8 

343.0 

625.5 

1009.2 

1486.5 

2378 

3481 

For  any  other  length  divide  240  by  the  given  length  expressed  in  diameters  and  multiply 
the  tabular  quantity  by  the  square  root  of  this  quotient,  which  will  give  the  flow  for  one  pound 
loss  of  pressure.  Conversely,  dividing  the  given  length  by  240  will  give  the  loss  of  pressure 
for  the  flow  given  in  the  table. 


638 


STEAM  POWER  PLANT  ENGINEERING 


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CO'-i 


PIPING  AND  PIPE  FITTINGS  -  639 

According  to  the  formulas  in  Group  I,  Table  84,  the  weights  dis- 
charged vary  as  ]  d5  -f-  ( 1  4-  —7-  J  f   and  the  equation  becomes 


(152) 


+  3.6 

From  (150)  and  (153)  we  see  that  the  values  of  2V x  are  practically 
the  same  for  either  equation  when  the  ratio  of  d  to  dt  is  small  and  that 
they  differ  widely  for  large  ratios.  For  example,  according  to  (150), 
5.7  eight-inch  pipes  are  equivalent  in  capacity  to  one  sixteen-inch  pipe, 
whereas  (152)  gives  6.15.  The  difference  is  negligible.  Again, 
according  to  (150),  180  two-inch  pipes  are  equivalent  in  capacity  to 
one  sixteen-inch  pipe,  whereas  (153)  gives  274.  The  difference  is  con- 
siderable. Equation  (153)  is  most  commonly  accepted  and  is  the  basis 
of  Table  87. 

338.  Friction  through  Valves  and  Fittings.  —  The  formulas  out- 
lined in  Table  83  are  strictly  applicable  only  to  well-lagged  pipes,  free 
from  bends  or  obstructions  of  any  kind  such  as  valves  or  fittings, 
which  greatly  increase  the  resistance  of  the  flow  of  steam.  If  these 
obstructions  must  be  considered,  it  is  customary  to  allow  for  them  by 
assuming  an  added  length  of  straight  pipe  equivalent  in  resistance 
to  the  various  fittings  and  bends.  Unfortunately,  the  few  tests  which 
have  been  made  for  the  purpose  of  determining  the  resistance  of  vari- 
ous pipe  fittings  give  discordant  results,  and  in  the  absence  of  more 
recent  data  the  rules  given  by  Briggs  ("Warming  Buildings  by  Steam") 
are  probably  as  accurate  as  any. 

According  to  Briggs,  the  length  of  pipe  in  inches  equivalent  to  the 
resistance  of  one  standard  90-degree  elbow  is 

L  =  76  d  +  (l  +  ^f]  (154) 

and  to  that  of  one  globe  valve 

L  =  114  d^-  (l  +  Y)'  (155) 

The  resistance  of  gate  valves  is  not  considered. 


640 


STEAM  POWER  PLANT  ENGINEERING 


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PIPING  AND  PIPE  FITTINGS 


641 


cc 

I 


642  STEAM  POWER  PLANT  ENGINEERING 

339.  Exhaust  Piping,  Condensing  Plants.  —  The  exhaust  piping  in 
condensing  plants  is  arranged  either  according  to  (1)  the  independent 
or  (2)   the  central  condensing  system.     In  the  former  each  engine  is 
provided  with  an  independent  condenser  and  air  pump.     In  case  the 
vacuum  "  drops  "  or  it  is  desired  to  operate  non-condensing,  the  steam 
is  discharged  through  a  branch  pipe  with  relief  valve  to  the  atmos- 
phere, Figs.  3  and  219.     When  there  are  a  number  of  engines  in  one 
installation  the  atmospheric  pipes  lead  to  a  common  free  exhaust  main, 
which,  on  account  of  its  great  size,  is  ordinarily  constructed  of  light- 
weight riveted  steel  pipe.     The  short  connection  between  engine  and 
condenser  is  usually  made  with  lap-welded  steel  pipe,  since  riveted 
joints  are  apt  to  leak,  due  to  the  engine  vibrations.     In  a  central  con- 
densing plant,  Fig.  226,  the  several  engines  exhaust  through  a  com- 
mon main  into  a  single  large  condenser.     An  atmospheric  relief  valve 
is  usually  provided  in  connection  with  the  condenser,   and  no  free 
exhaust  main  is  necessary.     Several  arrangements  of  condenser  piping 
are  illustrated  in  Figs.  219  to  228. 

340.  Exhaust    Piping,     Non-Condensing   Plant.  —  Webster   Vacuum 
System.     In  the  majority  of  non-condensing  plants  the  exhaust  steam 
is   used   for   heating   purposes.     One   of   the   best-known   systems   of 
exhaust  steam  heating,  in  which  the  back  pressure  on  the  engine  is 
reduced   by  circulating  below  atmospheric  pressure,  is  that   known  as 
the  Webster  combination   system.     The   general   arrangement   is  illus- 
trated in  Fig.  2  and  the  principles  of  operation  are  described  in  para- 
graph  3.      It   has   the   advantage   of   affording     (1)    minimum   back 
pressure  on  the  engine;     (2)    effective   and   continuous   drainage   of 
condensation    from    supply    pipes     and     radiators;      (3)  continuous 
removal  of   air  and  entrained  moisture  from  confined  spaces;    (4)  inde- 
pendent  regulation   of   temperature   in   each   radiator;     (5)  continu- 
ous return  of  condensation  to  the  boiler;    (6)  utilization  of  part  of 
the  exhaust  steam  for  preheating  the  feed  water;    and  (7)  automatic 
regulation.     Fig.  363  gives  a  diagrammatic  arrangement  of  the  piping 
and  appurtenances  in  a  typical  installation.     The  characteristic  feature 
of  this  system  is  the  automatic  outlet  valve  attached  to  each  part 
requiring  drainage,  which  permits  both  the  water  of  condensation  and 
the  non-condensable  gases  to  be  removed  continuously.     The  radiator 
temperature  may  be  regulated  by  varying  the  quantity  of  steam  sup- 
plied either  by  hand,  or  automatically  by  thermostatic  control.      The 
Webster  valve,  Fig.  364,  enables  the  vacuum  to  withdraw  the  water  of 
condensation  as  fast  as  it  is  formed  irrespective  of  the  pressure  in  the 
radiator,  hence  the  supply  may  be  throttled  to  such  an  extent  that  the 
temperature  in  the  radiator  is  practically  as  low  as  that  of  steam  cor- 


PIPING  AND  PIPE  FITTINGS 


643 


responding  to  the  pressure  in  the  vacuum  line.  The  small  annular 
space  between  the  inner  tube  of  the  float  F  and  the  guide  H  permits 
of  a  vacuum  in  the  body  of  the  valve.  When  the  water  from  the 
radiator  lifts  the  float  the  water  is  drawn  into  the  returns  pipe.  The 


OUTLET 

FIG.  364.    Webster  Air  Valve. 


FIG.    365.      Automatic  Vacuum 
Valve,  Illinois  Engineering  Co. 


valve  then  returns  to  its  seat  and  the  escape  of  steam  is  prevented 
except  such  as  finds  its  way  through  the  annular  space  around  the 
guide  stem  H.  An  improvement  on  this  valve  which  prevents  the 
escape  of  steam  is  illustrated  in  Fig.  365.  When  steam  is  admitted  to 
the  radiator  the  condensation  flows  into  the  valve,  righting  the  float  A 
and  sealing  the  outlet  B  against  the  passage  of  steam;  as  the  valve 
fills  with  water  the  buoyancy  of  the  float  raises  it  from  its  seat  and 
permits  the  water  to  be  drawn  out;  the  float  falls  and  reseats  on  the 
nipple  when  about  a  half-inch  of  water  remains  in  the  valve,  thus 
maintaining  a  water  seal. 

Screen  D  prevents  scale  and  dirt  from  entering  the  valve  proper. 
By-pass  H  is  for  emergency  use  in  draining  off  accumulated  water  in 
the  radiator  in  case  the  valve  becomes  stopped  up.  and  permits  the 
bonnet  to  be  removed  without  trouble  from  the  accumulated  water. 

341.  Exhaust  Piping,  Non-Condensing  Plants.  —  Paul  Heating  System. 
The  Paul  vacuum  system  differs  from  the  Webster  in  that  the  con- 
densation, and  the  air  and  non-condensable  gases  are  separately 
handled.  Referring  to  Fig.  366,  which  gives  a  diagrammatic  arrange- 
ment of  the  piping,  the  condensed  steam  gravitates  to  the  automatic 
return  tank  and  pump  and  is  pumped  either  directly  to  the  boiler  or 
through  the  heater  to  the  boiler.  Air  and  vapor  are  withdrawn  from 
the  upper  part  of  the  radiator  by  the  Paul  exhauster  or  ejector,  E, 
and  discharged  into  the  returns  tank,  which  is  vented  to  the  atmos- 
phere for  the  escape  of  the  non-condensable  gases.  The  exhauster 


644 


STEAM  POWER  PLANT  ENGINEERING 


PIPING  AND  PIPE  FITTINGS 


645 


receives  its  supply  of  steam  through  pipe  0,  Fig.  367,  which  shows  the 
general  arrangement  of  this  apparatus.  The  piping  is  in  duplicate  to 
guard  against  failure  to  operate.  The  suction  side  of  the  exhauster  is 
connected  with  the  air  pipes 
A,  A,  Fig.  366.  Fig.  368 
gives  a  section  through  the 
Paul  air  or  vacuum  valve 
which  prevents  steam  from 
blowing  into  the  air  pipes 
and  permits  only  air  to 
pass.  In  Fig.  366  the  heat- 
ing system  is  piped  on  what 
is  known  as  the  "  one-pipe 
down-feed  system;"  i.e., 
the  exhaust  steam  is  first 
conducted  to  a  distributing 
header  in  the  attic,  from 
which  the  various  supply 
pipes  are  led  to  the  radia- 
tors. The  water  of  conden- 
sation returns  through  these 
same  pipes  and  gravitates  STEAM  SUPPLy 

to  the  returns  pump.     Both  FIG.  367.     Paul  Exhauster. 

the  supply  steam  and  the 

condensation  flow  in  the  same  direction.  This  system  is  also  piped  on 
the  "  one-pipe  up-feed,"  the  "  two-pipe  up-feed,"  and  the  "  two-pipe 
down-feed  "  principle.  The  "  one-pipe  up-feed  "  differs  from  the  sys- 
tem just  described  in  that  the 
steam  flows  upward  through 
the  risers  and  does  away  with 
the  attic  piping.  The  returns, 
however,  flow  against  the  current 
of  steam  and  water  hammer 
is  more  likely  to  occur  than 
with  the  down-feed  system.  In 
the  two-pipe  systems  the  steam 
supply  pipes  or  risers  conduct 
steam  only,  and  the  returns 
carry  the  condensation.  The 
one-pipe  down-feed  is  cheaper  and  simpler  and  practically  as  efficient 
as  the  two-pipe  system  under  normal  conditions.  It  is  objection- 
able, however,  due  to  the  difficulty  of  draining  the  radiator  with 


COMPOSITION 


ROM  RADIATOR 


TO  EXHAUSTER 


FIG.  368.     Paul  Vacuum  Valve. 


646 


STEAM  POWER  PLANT  ENGINEERING 


closely  throttled  supply  valve,  since  the  velocity  of  the  entering 
steam  prevents  the  water  from  returning  through  the  same  orifice. 
342.  Automatic  Temperature  Control.  —  Experience  shows  that  a 
considerable  saving  in  fuel  may  be  effected  in  the  heating  plants  of  tall 
office  buildings  and  similar  plants  by  automatically  controlling  the 
temperature.  Hand-controlled  valves  are  usually  left  wide  open,  and 
when  the  room  becomes  too  hot  the  temperature  is  frequently  lowered 
by  opening  the  window,  resulting  in  a  waste  of  heat  which  may  be 
considerable  in  modern  buildings  with  hundreds  of  offices.  Many 
successful  methods  of  automatic  temperature  control  are  available, 
the  usual  system  consisting  of  thermostats  which  control  the  supply  of 
heat  by  means  of  diaphragm  valves,  the  latter  taking  the  place  of  the 
usual  radiator  supply  valve. 

Fig.  369  shows  a  Powers  thermostat.     The  expansible  disk  U  con- 
tains a  volatile  liquid  having  a  boiling  point  of   about  50  degrees  F. 

The  pressure  of  the  vapor  within  the  disk 
at  a  temperature  of  70  degrees  amounts 
to  six  pounds  to  the  square  inch,  and 
varies  with  every  change  of  temperature, 
causing  a  variation  in  the  thickness  of  the 
disk.  The  disk  is  attached  by  a  single 
screw  0  to  the  lever  Q,  which  rests  upon 
the  screw  F  as  a  fulcrum.  The  flat  spring  R 
holds  the  lever  and  disk  against  the  mova- 
ble flange  M.  Connecting  with  the  cham- 
ber N  are  two  air  passages  H  and  I.  The 
thermostat  is  attached  by  means  of  two 
screws  at  the  upper  end  to  a  wall  plate  per- 
manently secured  to  the  wall.  This  wall 
plate  has  ports  registering  with  H  and  7, 
one  for  supplying  air  under  pressure  and  the 
other  for  conducting  it  to  the  diaphragm 
motor  which  operates  the  valve  or  damper. 
Air  is  admitted  through  H  under  a  pressure 
of  about  fifteen  pounds  per  square  inch,  and 
its  passage  into  chamber  N  is  regulated  by 
the  valve  /,  which  is  normally  held  to  its 
seat  by  a  coil  spring  under  cap  P.  K  is  an 
elastic  diaphragm  carrying  the  flange  M,  with 
escape  valve  passage  covered  by  the  point  of  valve  L.  Valve  L  tends 
to  remain  open  by  reason  of  the  spring.  When  the  temperature  rises 
sufficiently  expansion  of  the  disk  U  first  causes  the  valve  to  seat,  its 


FIG.  369      Section  Through 
Powers  Thermostat. 


PIPING  AND  PIPE  FITTINGS 


647 


spring  being  weaker  than  that  above  valve  J.  If  the  expansive 
motion  is  continued,  valve  J  is  lifted  from  its  seat  and  compressed  air 
flows  into  chamber  N,  exerting  a  pressure  upon  the  elastic  diaphragm  K 
in  opposition  to  the  expansive  force  of  the  disk.  If  the  tempera- 
ture falls,  the  disk  contracts  and  the  overbalancing  air  pressure  in 
N  results  in  a  reverse  movement  of  the  flange  M,  permitting  the  escape 
valve  to  open  and  discharge  a  portion  of  the  air;  thus  the  air  pressure 
is  maintained  always  in  direct  propor- 
tion to  the  expansive  power  (and  tem- 
perature) of  the  disk  U.  The  passage  / 
communicates  with  a  diaphragm  valve, 
Fig.  370.  The  compressed  air  operates 
the  diaphragm  against  a  coiled  spring 
resistance,  so  that  the  movement  is 
proportional  to  the  air  pressure  and 
the  supply  of  steam  controlled  accord- 
ingly. The  adjusting  screw  G,  squared 
to  receive  a  key,  carries  an  indicator  by 
means  of  which  the  thermostat  can  be 
set  to  carry  any  desired  temperature 
within  its  range,  usually  from  60  to  80 
degrees.  In  changing  the  temperature 
adjustment  lever  Q  forces  the  disk  U 
closer  to  or  farther  from  the  flange  M. 

In   connecting   up    the    system    com- 
pressed air  is  carried  to  the  thermostat  and  diaphragm  valves,  from 
a  reservoir  through  small  concealed  pipes. 

In  the  indirect  system  of  heating  the  dampers  are  of  the  diaphragm 
type  and  the  method  of  regulation  is  the  same  as  with  the  direct  system. 

343.  Feed- Water  Piping.  —  The  simplest  arrangement  of  feed- 
water  piping  may  be  found  in  non-condensing  plants,  in  which  the 
feed  water  is  obtained  under  a  slight  head,  such  as  is  afforded  by  the 
average  city  supply,  and  is  heated  in  an  open  heater  by  the  exhaust 
steam  from  the  engine  to  a  temperature  varying  from  180  to  210 
degrees  F.  The  hot  feed  water  gravitates  from  the  heater  to  the 
pump  and  then  is  forced  to  the  boiler,  or  to  the  economizer  if  one  is 
used.  If  a  meter  is  used  it  is  generally  placed  on  the  discharge  side 
of  the  pump,  and  should  be  by-passed  to  permit  it  to  be  cut  out  for 
repairs.  (Fig.  371.)  Plants  operating  continuously  should  have  feed 
pumps  in  duplicate.  In  some  cases  the  returns  from  the  heating 
system  gravitate  to  the  heater  and  only  enough  cold  water  is  added  to 
make  up  the  loss  from  leakage,  etc.  In  other  cases  the  returns  gravi- 


FIG.  370. 


A    Typical 
Valve. 


Diaphragm 


648 


STEAM   POWER  PLANT  ENGINEERING 


tate  to  a  special  "  returns  tank,"  from  which  they  are  pumped  directly 
to  the  boiler  without  further  heating.  Occasionally  a  live-steam 
purifier  is  used,  especially  if  the  water  contains  a  large  percentage  of 


O-Q 


i> 


CX£) 


LER                                                 I                BOILER                                            _L_  BOILER 
.N.CCTOR       '     MA.N ' *T*1        ""^"H}*-; 


COLD   WATER    SUPPL1 


FIG.  371.     Feed  Water  Piping;  No^Condensing  Plant. 

calcium  sulphate.  The  feed  is  then  subjected  to  boiler  pressure  and 
temperature  and  the  greater  part  of  the  impurity  precipitated  before  it 
enters  the  boiler.  Closed  heaters  are  often  used  in  place  of  open 
heaters.  When  the  supply  is  not  under  head  a  closed  heater  is  usually 
preferred  and  is  placed  between  the  pump  discharge  and  the  feed 
main. 

In  condensing  plants  the  feed  piping   is  similar  to   that  in  non- 
condensing  plants,  except  that  if  exhaust  steam  is  used  for  heating 


FIG.  372.     Feed  Water  Piping;  Condensing  Plant. 

purposes  it  is  supplied  by  the  auxiliaries,  such  as  feed  pumps,  stoker 
engines,  condenser  engines,  and  other  steam-using  appliances. 

In  plants  having  a  number  of  boilers  it  is  customary  to  run  a  feed 
main  or  header  the  full  length  of  the  boiler  room  and  connect  it  to 


PIPING  AND  PIPE  FITTINGS 


649 


each  boiler  by  a  branch  pipe.  This  main  may  be  a  simple  header  or 
in  duplicate  or  of  the  "  loop  "  or  "  ring  "  type.  Horizontal  tubular 
boilers  are  frequently  arranged  in  one  battery  with  the  feed  main  run 
along  the  fronts  of  the  boilers  just  above  the  fire  doors.  Water-tube 
boilers  are  generally  set  in  a  battery,  and  as  the  arrangement  above 
would  block  the  passageway  between  the  batteries,  the  main  is  run 
either  above  or  under  the  settings,  the  former  being  the  more  common. 
Where  a  single  header  is  used,  the  feed  pumps  are  sometimes  placed  so 
as  to  feed  into  opposite  ends  of  the  main,  which  is  then  cut  into 
sections  by  valves.  Another  arrangement  is  to  place  the  pumps  so  as 
to  feed  into  the  middle  of  the  header.  With  the  loop  arrangement  the 
main  is  ordinarily  cut  into  sections  by  valves  so  that  the  water  may  be 
sent  either  way  from  the  pumps  and  any  defective  section  cut  out. 
With  duplicate  mains  a  common  arrangement  is  to  place  one  main 
along  the  front  of  the  boiler  and  the  other  at  the  rear  or  both  over- 
head as  in  Fig.  359.  Sometimes  one  main  is  placed  in  the  passageway 
below  the  boiler  setting  and  the  other  on  top. 

Standard  wrought-iron  pipe  is  usually  used  for  pressures  under  100 
pounds  and  extra  heavy  pipe  for  greater  pressures.  The  pipes  and 
fittings  from  boiler  to  main  are  frequently  of  brass,  and  preferably  so, 
since  brass  withstands  corrosive  action  much  better  than  iron  or  steel. 
Flanged  joints  should  be  used  in  all  cases,  since  the  pockets  formed  by 
the  ordinary  screwed  joints  hasten  corrosion  at  those  points.  (Power, 
June,  1902,  p.  4.) 

Fig.  373,  A  to  E,  illustrates  the  various  combinations  of  check  valve, 
stop  valves",  and  regulating  valve  in  steam  boiler  practice.  The 


FIG.  373.    Different  Arrangements  of  Valves  in  Feed  Water  Branch  Pipes. 

simplest  arrangement  and  one  sometimes  used  in  plants  operating 
intermittently  is  shown  in  A .  Here  there  are  but  two  valves  between 
the  boiler  and  the  main,  the  check  being  nearest  the  boiler  and  the  stop 
valve  at  the  main.  The  stop  valve  performs  both  the  function  of 
cutting  out  the  boiler  and  of  regulating  the  water  supply.  This 


650  STEAM  POWER  PLANT  ENGINEERING 

arrangement  is  not  recommended,  as  any  sticking  or  excessive  leaking 
of  the  check  valve  will  necessitate  shutting  down  the  boiler.  B  shows 
the  most  common  arrangement.  Here  the  check  valve  is  placed 
between  the  regulating  valve  and  a  stop  valve  as  indicated.  This 
permits  a  disabled  check  to  be  easily  removed  while  pressure  is  on  the 
boiler  and  the  main.  E  shows  an  arrangement  whereby  both  check 
and  regulating  valve  may  be  removed,  and  is  particularly  adapted  to 
boilers  operating  continuously  where  the  regulating  valve  is  subjected 
to  severe  usage.  In  this  case  the  stop  valves  are  run  wide  open  and 
are  subjected  to  no  wear.  The  regulating  valve  most  highly  recom- 
mended is  a  self-packing  brass  globe  valve  with  regrinding  disk.  The 
check  valve  is  ordinarily  of  the  swing  check  pattern  with  regrinding 
disk,  Fig.  384  (C).  Modern  practice  recommends  an  automatic  water 
relief  valve  in  the  discharge  pipe  immediately  adj  acent  to  each  pump  to 
prevent  excessive  pressure  in  case  a  valve  is  accidentally  closed  in 
by-passing  or  in  changing  over. 

344.  Flow  of  Water  through  Orifices,  Nozzles,  and  Pipes.  —  Ber- 
noulli's theorem  is  the  rational  basis  of  most  empirical  formulas  for 
the  steady  flow  of  a  fluid  from  an  up-stream  position  n  to  a  down- 
stream position  m,  thus  ("Mechanics  of  Engineering/'  Church,  p.  706): 

p  y  2  p  ym  fall  losses  of  head  \ 

"T*  +  o  +  Z*»  =  +  "7T"  +  Z»  -  i  occurring  between  V>  (156) 

y  2°  ?  2<>  Uandm  J 

in  which 

V  =  velocity  in  feet  per  second  at  the  point  considered. 
P  =  pressure  in  pounds  per  square  foot. 
Z  =  potential  head  in  feet  of  the  fluid. 

y  =  density  of  the  fluid,  pounds  per  cubic  foot. 

g  =  acceleration  of  gravity. 

Each  loss  of  head  will  be  of  the  form    K—,  in  which  K  is  the 

2g 

coefficient  of  resistance  to  be  determined  experimentally.  The  loss  of 
head  due  to  skin  friction  is  expressed  : 


in  which 

/  =  the  coefficient  of  friction  of  the  fluid  in  the  pipe. 
I  =  length  of  the  pipe  in  feet. 
d  =  diameter  of  the  pipe  in  feet. 
Other  notations  as  in  (156). 


PIPING  AND  PIPE  FITTINGS  651 

Discharge  from  a  circular  vertical  orifice  with  sharp  corners: 

Q  =  CA  Vzljh,  (158) 

in  which 

Q  =  cubic  feet  per  second. 

C  —  coefficient,  varying  from  0.59  to  0.65  (Merriman,  Treatise  on 

Hydraulics,  p.  118). 
A  =  area  of  the  orifice,  square  feet. 
h  =  head  of  water  in  feet. 
g  =  acceleration  of  gravity  =  32.2. 

Discharge  from  short  cylindrical  nozzles  three  diameters  in  length,  with 
rounded  entrance  ("Mechanics  of  Engineering,"  Church,  p.  690): 

Q  =  0.815  A  V2~gh.  (159) 

Discharge  from  short  nozzles  with  well-rounded  corners  and  conical 
convergent  tubes,  angle  of  convergence  13£  degrees  (Church,  p.  693): 

Q  =  0.94  A  V^Tgh.  (160) 

Discharge  from    cylindrical    pipe    under    500    diameters    in    length 
(Church,  p.  712) : 


9  "<*v(i  + »*  +  •-' 


in  which 

f  =  coefficient  of  friction. 

Other  notations  as  above. 

/  varies  with  the  nature  of  the  inside  surface,  the  diameter  of  the 
pipe,  and  the  velocity  of  flow. 

Discharge  through  very  long  cylindrical  pipes  ("  Mechanics  of  Engi- 
neering," Church,  p.  715): 


Q  =  3.15  y         .  (162) 

Loss  of  head  due  to  friction  in  water  pipes.      Weisbach's  formula  is 
as  follows: 

0«3> 


in  which 

H  =  friction  head  in  feet. 

V  =  velocity  in  feet  per  second. 

L  =  length  of  pipe  in  feet. 

d  =  diameter  of  pipe  in  inches. 


652 


STEAM  POWER  PLANT  ENGINEERING 


TABLE    OF   THE    COEFFICIENT  /  FOR    FRICTION    OF   WATER    IN    CLEAN 

IRON    PIPES. 

(Abridged  from  Fanning.) 


Velocity  in 

Diam. 

Diam. 
=  1  in. 

Diam. 
=  2  in. 

Diam. 
=  3  in. 

Diam. 
=  4  in. 

Diam. 
=  6  in. 

Diam. 

=  8  in. 

Ft.  per  Sec. 

=  .0417  ft. 

=  .0834  ft. 

=  .1667  ft. 

=  .25  ft. 

=  .333  ft. 

=  .50  ft. 

=  .667  ft. 

0.1 

.0150 

.0119 

.00870 

.  00800 

.00763 

.00730 

.00704 

0.3 

.0137 

.0113 

850 

784 

750 

720 

693 

0.6 

.0124 

.0104 

822 

767 

732 

702 

677 

1.0 

.0110 

.00950 

790 

743 

712 

684 

659 

1.5 

.00959 

.00868 

.00757 

.00720 

.00693 

.00662 

.00640 

2.0 

.00862 

810 

731 

700 

678 

648 

624 

2.5 

795 

768 

710 

683 

662 

634 

611 

3.0 

.00753 

.00734 

.00692 

.00670 

.00650 

.00623 

.00600 

4.0 

722 

702 

671 

651 

631 

607 

586 

6.0 

689 

670 

640 

622 

605 

582 

562 

8.0 

663 

646 

618 

600 

587 

562 

544 

12.0 

630 

614 

590 

582 

560 

540 

522 

16.0 

.00618 

.00600 

.00581 

.00570 

.00552 

.00530 

.00513 

20.0 

615 

598 

579 

566 

549 

525 

508 

Velocity  in 

Diam. 

=  10  in. 

Diam. 
=  12  in. 

Diam. 
=  16  in. 

Diam. 
=  20  in. 

Diam. 
=  30  in. 

Diam. 
=  40  in. 

Diam. 
=  60  in. 

Ft.  per  Sec. 

=  .833  ft. 

=  1.00  ft. 

=  1.333  ft. 

=  1.667  ft. 

=  2.50  ft. 

=  3.  333  ft. 

=  5.  ft. 

0.1 

.00684 

.00669 

.00623 

0  3 

673 

657 

614 

.00578 

0^6 

659 

642 

603 

567 

.00504 

.00434 

.00357 

1.0 

643 

624 

588 

555 

492 

428 

353 

1.5 

.00625 

.00607 

.00572 

.00542 

.00482 

.00421 

.00349 

2.0 

609 

593 

559 

529 

470 

416 

346 

2.5 

596 

581 

548 

518 

460 

410 

342 

3.0 

.00584 

.00570 

.00538 

.00509 

.00452 

.00407 

.00339 

4.0 

568 

553 

524 

498 

441 

400 

333 

6.0 

548 

534 

507 

482 

430 

391 

324 

8.0 

532 

520 

491 

470 

422 

384 

320 

12.0 

512 

500 

478 

457 

412 

377 

.00313 

i  ""•  16.0 

.00502 

.00491 

.00470 

.00450 

.00406 

.00370 

20.0 

498 

485 

William  Cox  (American  Machinist,  Dec.  28,   1893)  gives   a  simple 
formula  which  gives  almost  identical  results: 

„       (4  V»  +  57-2)L 


1200  d 
Notations  as  in  (163). 


(164) 


Loss  of  head  due  to  friction  of  fittings.     Formulas  (161)  to  (164)  are 
based  on  the  flow  of  water  through  clean  straight  cylindrical  pipes. 


PIPING  AND  PIPE  FITTINGS  653 

Where  there  are  bends,   valves,   or  fittings  in  the  line  the  flow  is 
decreased  on  account  of  the  additional  resistance. 

These  frictional  losses  are  conveniently  expressed  in  feet  of  water, 
thus: 

H  =  C  p  •  (165) 

C  having  the  following  values: 

Angles.  Class  of  Valve. 

45  degrees.     90  degrees.  Gate.     Globe.     Angle, 

C  0.182  0.98  0.182       1.91         2.94 

Example:  Determine  the  pressure  necessary  to  deliver  200  gallons 
of  water  per  minute  through  a  4-inch  iron  pipe  line  400  feet  long,  fitted 
with  four  right-angle  elbows  and  two  globe  valves.  The  water  is  to 
be  discharged  into  an  open  tank. 

A  flow  of  200  gallons  per  minute  gives  a  velocity  of 

900  V   1  44 

-        *        -  =  5  feet  per  second  (7.48  =  number  of  gallons  per 
7.48  X  60  X  12.72 

cubic  foot,  and  12.72  =  internal  area  of  the  pipe,  square  inches). 
From  the  preceding  table,  /  =  0.618  for  7  =  5. 
From  (165), 

OK 

Resistance  head  of  4  elbows  =  0.98  X  -=^-  X  4  =  1.52  feet. 

64.4 

Resistance  head  of  2  globe  valves: 

OK 

1.91  X  ~  X  2  =  1.48  feet. 
64.4 

Resistance  head  of  all  fittings: 

1.52+  1.48  =  3  feet. 
Substitute  7  =  5,  L=  400,  and  d  =  4  in  (164). 


V          1200  X  4 
=  10.25  feet,  resistance  head  of  the  pipe. 

Total  resistance  head  =  10.25  +  3  =  13.25  feet  of  water,  or  5.75 
pounds  per  square  inch. 

Example:  How  many  gallons  of  water  will  be  discharged  per  minute 
through  above  line  with  initial  pressure  of  100  pounds  per  square  inch, 
and  what  will  be  the  pressure  at  the  discharge  end? 

Since  /  depends  upon  the  unknown  7,  we  may  put  /  =  0.006  for  a 
first  approximation  and  solve  for  7;  then  take  a  new  value  of  /  and 
substitute  again,  and  so  on. 


654 


STEAM  POWER  PLANT  ENGINEERING 


Substitute /=  0.006,  d=  -2-  ,h=  100X2.3  =  230,  and Z  =  400 in (162): 


Q  =  3.15 


I.335 


230 


1.006  X  400 

=  1.95   cubic   feet   per  second,    corresponding  to   a 
velocity  of  22  feet  per  second. 

From  the  preceding  table, 

/  =  0.00548  (by  interpolation)  for  V  =  22  feet  per  second. 

From  (165)  the  friction  of  4  elbows  and  2  globe  valves  is  found  to 
be  58  feet  for  V  =  22. 

From  (164)  a  resistance  head  of  58  feet  of  water  for  V  =  22  is 
found  to  be  equivalent  to  136  feet  of  straight  pipe,  thus: 
X  222  X  5  X  22-2 


58 


1200  X  4 


L. 


L  =  136. 

Substitute/  =  0.0548, 1  =  400  +  136  =  536  in  (162): 
L335  X  230 


).0058  X  536 
=  1.74  cubic  feet  per  second,  corresponding  to  a  velocity 

of  19.3  feet  per  second. 
=  780  gallons  per  minute. 

If  greater  accuracy  is  necessary  determine  /  and  L  for   V  =  19.3 
and  proceed  as  above. 

The  total  friction  head  may  be  determined  from  (164)  thus: 
X  19.32  +  5  X  19.3-2 

1200  X  4 
=  177  feet  of  water. 
=  77  pounds  per  square  inch. 

The  pressure  at  the  discharge  end  will  be 

100  —  77  =  23  pounds  per  square  inch. 

Average  power  plant  practice  gives  the  following  maximum  veloci- 
ties of  flow  in  water  pipes : 


Size  of  Pipe  in 
Inches. 

Velocity,  Feet  per 
Minute. 

Size  of  Pipe  in 
Inches. 

Velocity,  Feet  per 
Minute. 

i*M 

I  tali 

1J  to  3 

50 
100 
200 

3  to  6 
Over  6 

250 
300-400 

PIPING  AND  PIPE  FITTINGS 


655 


345.  Stop  Valves.  —  The  valves  used  to  control  and  regulate  the 
'flow  of  fluids  are  the  most  important  element  in  any  piping  system. 
A  good  valve  should  have  sufficient  weight  of  metal  to  prevent 
distortion  under  varying  temperature  and  pressure,  or  under  strains 
due  to  connection  with  the  piping;  the  seats  should  be  easily  repaired 
or  renewed;  there  should  be  no  pockets  or  projections  for  the  accu- 
mulation of  dirt  and  scale,  and  the  valve  stem  should  permit  of  easy 
and  efficient  packing.  Stop  valves  are  made  in  such  a  variety  of 
designs  that  a  brief  description  will  be  given  of  only  a  few  of  the  best- 
known  types. 

Fig.  374  shows  a  section  of  an  ordinary  globe  valve,  so  called  because 
of  the  globular  form  of  the  casing.  This  type  of  valve  is  the  most 


c 


FIG.  374.    A   Typical  Globe    Valve, 
Screw-Top,  inside  Screw. 


FIG.  375.     A  Typical  Globe  Valve, 
Bolt-Top,  outside  Screw. 


common  in  use.  Globe  valves  are  designated  as  (1)  inside  screw  and 
(2)  outside  screw,  according  as  the  screw  portion  of  the  stem  is  inside 
the  casting,  Fig.  374,  or  outside,  Fig.  375.  The  top  or  bonnet  may  be 
screwed  into  the  body  of  the  valve,  Fig.  374,  or  bolted,  Fig.  375.  The 
smaller  sizes,  three  inches  and  under,  are  usually  of  the  screw-top  type 
and  the  larger  of  the  bolt-top  type.  Valves  with  outside  yoke  and 
screw  are  preferable  to  the  other  in  that  they  show  at  a  glance  whether 
the  valve  is  open  or  closed,  an  advantage  in  changing  from  one  section 
to  another.  The  disks  are  made  in  a  variety  of  forms,  the  material 


656 


STEAM  POWER  PLANT  ENGINEERING 


depending  upon  the  nature  of  the  fluid  to  be  controlled.  Thus,  for 
cold  water,  hard  rubber  composition  gives  good  results;  for  hot  water* 
and  low-pressure  steam,  Babbitt  metal;  for  high-pressure  steam, 
copper  or  bronze;  and  for  highly  superheated  steam,  nickel.  The 
valve  bodies  are  of  brass  for  sizes  under  three  inches,  cast  iron  for  the 
larger  sizes  and  ordinary  pressures  and  temperatures,  and  cast  steel  or 
semi-steel  for  high  temperatures  and  pressures.  Globe  valves  should 
always  be  set  to  close  against  the  pressure,  otherwise  they  could  not 
be  opened  if  the  valves  should  become  detached  from  the  stem.  Globe 
valves  should  never  be  placed  in  a  horizontal  steam  return  pipe  with 


FIG.  376.  A  Typical  Gate 
Valve,  Solid-Wedge,  Screw- 
Top,  outside  Screw. 


FIG.  377.  A  Typical  Gate  Valve, 
Solid- Wedge,  Bolt-Top,  inside 
Screw. 


the  stem  vertical,  because  the  condensation  will  fill  the  pipe  about 
half  full  before  it  can  flow  through  the  valve.  Globe  valves  that  are 
open  all  the  time  are  preferably  designed  with  a  self-packing  spindle,  as 
in  Fig.  375,  in  which  the  top  of  shoulder  C  can  be  drawn  tightly 
against  the  under  surface  of  bonnet  S,  thus  preventing  steam  from 
leaking  past  the  screw  threads  while  the  spindle  is  being  packed. 

Figs.  376  to  379  show  different  types  of  gate  or  straight  way  valves. 
These  valves  offer  little  resistance  to  the  flow  of  steam  or  liquid  passing 
through  them,  and  are  generally  used  in  the  best  class  of  work. 
Fig.  376  shows  a  section  through  a  solid-wedge  gate  valve  with  outside 


PIPING  AND  PIPE  FITTINGS 


657 


screw  and  yoke.  This  form  of  outside  screw  and  yoke  with  stem 
protruding  beyond  the  hand  wheel  is  a  perfect  indicator  to  show 
whether  the  valve  is  open  or  shut,  as  the  hand  wheel  is  stationary 
and  the  spindle  rises  in  direct  proportion  to  the  amount  the  valve  is 
opened.  For  these  reasons  outside  screw  valves  are  preferable  for 
high-pressure  work  and  especially  for  the  larger  sizes.  The  seats  are 
made  solid,  or  removable,  and  of  various  materials  for  different  pres- 
sures and  temperatures.  Fig.  378  shows  a  section  through  a  split- 
wedge  gate  valve  with  parallel  faces  and  seats.  For  the  sake  of  illus- 


FIG.  378.    A  Typical  Gate  Valve,  Split- 
Wedge,  Bolt-Top,  inside  Screw. 


FIG.  379.    Ludlow  Angle  Valve,  Gate 
Pattern. 


tration  this  valve  is  fitted  with  inside  screw.  In  this  design  the 
spindle  remains  stationary  so  far  as  any  vertical  movement  is  con- 
cerned, and  the  gate  or  plug  being  attached  to  it  by  means  of  a 
threaded  nut  rises  into  the  bonnet  when  the  spindle  is  revolved. 
It  is  impossible  to  tell  by  its  appearance  whether  this  form  of  valve 
is  open  or  closed.  Valves  with  inside  screw  are  adapted  to  situa- 
tions where  there  is  considerable  dirt  and  grit,  since  the  screw  is 
inclosed  and  protected  and  excessive  wear  is  thus  avoided.  Gate 
valves  with  split  gates  are  more  flexible  than  those  with  solid  gates, 
and  hence  are  less  likely  to  leak.  Fig.  379  shows  the  application 


658 


STEAM  POWER  PLANT  ENGINEERING 


of  the  gate  system  to  an  angle  valve.  All  high-pressure  valves  above 
8  inches  in  diameter  should  be  provided  with  a  small  by-pass  valve, 
as  the  pressure  exerted  against  the  disk  or  gate  is  very  great  when  the 
valve  is  closed  and  the  force  required  to  move  it  is  considerable.  The 
by-pass  valve  also  facilitates  "  warming  up  "  the  section  to  be  cut  in 
and  is  more  readily  operated  than  the  main  valve. 

346.  Automatic    Non-Return    Valves.  —  Fig.    380    shows    a   section 
through  an  automatic  non-return  valve  as  applied  to  the  nozzle  of  a 

steam  boiler.  As  will  be  seen  from  the 
illustration  it  amounts  to  practically  a 
large  check  valve  with  cushioned  disk. 
The  object  of  this  device  is  the  equali- 
zation of  pressure  between  the  different 
units  of  the  battery,  the  valve  remaining 
closed  as  long  as  the  individual  boiler 
pressure  is  lower  than  that  of  the  header. 
In  case  a  tube  blows  out  the  valve  closes 
automatically,  owing  to  the  reduction  of 
pressure  and  prevents  the  header  steam 
from  entering  the  boiler.  It  acts  also 
as  a  safety  stop  to  prevent  steam  being 
turned  into  a  cold  boiler  while  men  are 
working  inside,  because  it  cannot  be 
opened  when  there  is  pressure  on  the 

header    side    only.       To     be     successful, 
FIG.  380.    Andereon   Non-Return    guch    ft  yalve    gh()uld    not   Qpen  untn    the 

pressure  in  the  boiler  is  equal  to  that  in 

the  header;  it  should  not  stick  and  become  inoperative  nor  chatter 
and  hammer  while  performing  its  work.  Referring  to  Fig.  380,  tail 
rod  E  insures  alignment  and  hence  prevents  sticking;  steam  space  C 
acts  as  a  dashpot  to  prevent  hammering  of  the  valve  as  it  rises,  and 
steam  space  D  acts  as  a  cushion  and  prevents  hammering  at  closing. 
Lip  F  is  made  to  enter  the  opening  in  the  seat  and  reduce  wire  draw- 
ing across  the  seat.  Fig.  358  shows  the  installation  of  a  number  of 
non-return  valves  at  the  Yonkers  power  house  of  the  New  York  Central 
Railway  Company. 

347.  Emergency  Valves.  —  In  large  power  plants  it  is  customary  to 
protect  the  various  divisions  of  the  steam  piping  by  emergency  valves 
which  may  be  closed  by  suitable  means  at  any  reasonable  distance 
from  the  valve.     The  simplest  form  of  emergency  stop  is  a  weighted 
"  butterfly  "  valve,  which  is   to   all  intents  and  purposes  a  weighted 
check  as  illustrated  in  Fig.  385  (F) .    The  weight  when  supported,  say 


PIPING  AND  PIPE  FITTINGS 


659 


by  a  cord  and  pulley,  holds  the  valve  open;  when  the  cord  is  cut  or 
released  the  weight  drops  and  forces  the  valve  shut.  The  cord  may 
lead  to  any  convenient  and  safe  distance  from  the  valve.  In  applying 
this  system  of  control  to  steam  engines  the  valve  is  placed  in  the 
steam  pipe  just  above  the  throttle  and  the  weight  held  up  by  a  lever 
controlled  by  the  main  governor  or  preferably  by  a  separate  gov- 
ernor. Should  the  engine  exceed  a  certain  speed,  as  in  case  of  accident 
to  the  regular  governor,  the  lever  supporting  the  weight  is  tripped 
by  the  emergency  governor  and  the  valve  is  closed  automatically. 
For  high  pressures  a  rotating  plug  valve  or  cock  is  preferred  to  the 
butterfly  type,  since  it  is  balanced  in  all  positions.  Gate  and  globe 
valves  may  be  converted  into  emergency  valves  by 
having  the  stems  mechanically  operated  by  electric 
motors,  hydraulic  pistons,  and  the  like.  Fig.  381 
shows  a  section  through  a  Crane  hydraulically  oper- 
ated emergency  gate  valve. 


FIG.  381.  Crane 
Emergency  Valve, 
Hydraulic. 


FIG.  382.     Anderson  Triple- 
Duty  Emergency  Valve. 


FIG.  383.  Pilot  Valve  for 
Anderson  Triple-Duty 
Emergency  Valve. 


Fig.  382  shows  a  partial  section  through  an  "  Anderson  triple- 
duty  "  emergency  valve,  and  Fig.  383  a  section  through  the  pilot 
valve.  A  steam  connection  from  the  main  line  to  the  top  of  a  copper 
diaphragm  holds  the  pilot  valve  closed  because  of  the  large  area 
above  the  diaphragm.  A  steam  pipe  connection  from  underneath 
the  emergency  piston  of  the  triple-acting  valve  also  leads  to  the  pilot 


660  STEAM  POWER  PLANT  ENGINEERING 

valve.  In  case  a  break  occurs  in  the  main  steam  line  or  branches, 
the  pressure  is  removed  from  the  top  of  the  pilot  valve,  causing  it  to 
open/  thus  exhausting  the  pressure  from  beneath  the  emergency  piston 
in  the  triple-acting  valve.  The  boiler  pressure  on  top  of  the  emer- 
gency piston  causes  the  valve  to  close.  Pilot  valves  may  be  located 
at  any  desirable  places,  thus  affording  control  from  different  points. 

In  the  "  Locke  automatic  engine  stop  system  "  the  stop  valve  is 
operated  by  an  electric  motor  which  is  controlled  by  contact  points 
operated  by  a  speed-limit  device.  (See  Power,  August,  1907,  p.  471, 
for  a  detailed  description.) 

348.  Check  Valves.  —  Fig.  384,  A  to  D,  illustrate  the  different 
types  of  check  valves  in  most  common  use.  A  is  a  ball  check,  B  a  cup 


-(A)  (B)  (c)  (Oj 

FIG.  384.    Types  of  Check  Valves. 

or  disk  check,  C  a  swing  check,  and  D  a  weighted  check.  Occasionally 
the  valve  body  is  fitted  with  a  valve  stem  and  handle  for  holding  the 
disk  against  its  seat,  in  which  it  is  designated  as  a  stop  check.  In  A 
and  B  the  valve  seat  is  parallel  to  the  direction  of  flow  and  the  valve 
is  held  in  place  by  its  own  weight  and  by  the  pressure  of  the  fluid  in 
case  of  reverse  flow.  In  the  swing  check  the  seat  is  at  an  angle  of 
about  45  degrees  to  the  direction  of  flow.  The  latter  construction  is 
preferred  as  it  offers  less  resistance  to  flow  and  there  is  less  tendency 
for  impurities  to  lodge  on  the  valve  seat.  By  extending  the  hinge  of 
the  swing  through  the  body  of  the  valve,  a  lever  and  weight  may  be 
attached  as  in  D  and  the  check  will  not  open  except  at  a  pressure 
corresponding  to  the  resistance  of  the  weight.  It  thus  acts  as  a  relief 
valve  and  at  the  same  time  prevents  a  reversal  of  flow.  Stop  checks 
are  usually  inserted  in  boiler  feed  lines  close  to  the  boiler,  and  when 
locked,  act  as  any  ordinary  stop  valve  and  permit  the  piping  to  be  dis- 
mantled or  the  regulating  valve  to  be  reground  without  lowering  the 
pressure  on  the  boiler.  Since  the  wear  on  check  valves  is  excessive 
and  necessitates  frequent  regrinding  they  are  often  mounted  with 
regrinding  disks,  Fig.  384  (C),  which  may  be  "  ground  "  against  the  seat 
without  removing  the  valve  from  the  line. 


PIPING  AND  PIPE  FITTINGS 


661 


349.  Blow-off  Cocks  and  Valves.  —  The  requirements  of  a  good 
blow-off  valve  are  that  it  shall  furnish  a  free  passage  for  scale  and 
sediment,  that  it  shall  close  tightly  so  as  not  to  leak,  and  that  it  shall 
open  easily  without  sticking  or  cutting.  On  account  of  the  rather 
severe  service  to  which  such  valves  are  subjected,  they  are  made 
very  heavy,  with  renewable  wearing  parts. 

Fig.  385  gives  a  sectional  view  of  a  Crane  ferrosteel  valve.  The 
bonnet  is  easily  taken  off  and  the  disk  removed  to  be  refaced  or 
replaced  by  a  new  one.  The  old  disk  is  repaired  by  pouring  in  a  hard 
Babbitt  metal  and  facing  it  off  flush.  The  seats  are  of  brass  and  oval 
on  top  to  prevent  scale  lodging  between  them  and  the  disk,  and  are  so 
made  that  they  may  be  removed;  but  it  has  been  found  in  practice 
that  there  is  not  much  cutting  of  the  seat,  the  damage  usually  being 
confined  to  the  softer  Babbitt  metal  which  faces  the  disk. 


FIG.  385.    Crane  Ferrosteel    Fia.  387. 
Blow-off  Valve. 


A  Typical  Blow-off 
Cock. 


FIG.  386.    Faber  Blow-off 
Valve. 


Fig.  386  gives  a  sectional  view  of  a  Faber  valve.  When  the  disk, 
which  makes  a  snug  fit  in  the  body  of  the  valve,  is  in  the  position 
shown,  the  boiler  discharge  is  practically  shut  off  and  any  sediment 
lying  on  the  seat  is  cleaned  off  by  a  jet  of  steam  or  water. 

Fig.  387  shows  a  section  through  a  typical  blow-off  cock  of  the 
straightway  taper  plug  pattern  with  self-locking  cam.  Plug  cocks  are 
often  used  instead  of  valves  on  the  blow-off  piping. 

Current  practice  recommends  the  use  of  two  valves,  or  rather  one 
valve  and  one  cock,  in  the  blow-off  line  of  each  boiler.  In  most  of 


662 


STEAM  POWER  PLANT  ENGINEERING 


PIPING  AND  PIPE  FITTINGS 


663 


the  large  stations  a  blow-off  valve  and  a  blow-off  cock  are  installed  as 
indicated  in  Fig.  388.  The  number  and  size  of  blow-off  cocks  are 
usually  specified  by  city  or  state  legislation. 

350.  Safety  Valves.  —  Fig.  389  shows  a  section  through  the  simplest 
form  of  safety  valve.  The  valve  is  held  on  its  seat  against  the  boiler 
pressure  by  a  cast-iron  weight  as 
indicated.  This  type  has  the  ad- 
vantage of  great  simplicity,  and 
can  be  least  affected  by  tampering, 
since  it  requires  so  much  weight 
that  any  additional  amount  which 
would  seriously  overload  it  can  be 
quickly  detected.  For  high  pres- 
sure and  large  sizes  of  boiler  this 
class  of  valve  is  entirely  too  cum- 
bersome. 

Fig.   390    shows    the    general    de-     FIG.  389.    "  Dead-weight "  Safety  Valve. 

tails  of  the  common  lever  safety  valve. 

The  valve  is  held  against  its  seat  by  a  loaded  lever,  thereby  enabling 
the  use  of  a  much  smaller  weight  than  the  "  dead  weight  "  type,  since 
the  resistance  is  multiplied  by  the  ratio  of  the  long  arm  of  the  lever  to 


TO     so     »o    loo    iio    120) 


FIG.  390.     Common  Lever  Safety  Valve. 

the  short  one.    The  proper  position  of  the  weight  is  determined  by 
simple  proportion. 

Fig.  391  shows  a  section  through  a  typical  pop  safety  valve  in  which 
the  boiler  pressure  is  resisted  by  a  spring.  This  type  of  valve  has 
practically  supplanted  all  other  forms.  The  boiler  pressure  acting  upon 
the  under  side  of  valve  V  is  resisted  by  the  tension  in  spring  S.  As 
soon  as  the  boiler  pressure  exceeds  the  resistance  of  the  spring  the 
valve  lifts  from  its  seat  and  the  steam  escapes  through  opening  0. 
The  static  pressure  of  the  steam  plus  the  force  of  its  reaction  in  being 
deflected  from  the  surface  A  holds  the  valve  open  until  the  pressure  in 


664 


STEAM  POWER  PLANT  ENGINEERING 


the  boiler  drops  about  5  pounds  below  that  at  which  the  valve  is  lifted. 
The  additional  area  of  valve  exposed  to  pressure  when  the  valve  lifts 

causes  it  to  open  with  a  sudden  motion 
which  has  given  it  its  name,  and  it  also 
closes  suddenly  when  the  pressure  has 
fallen.  These  valves  are  arranged  so 
that  the  spring  tension  may  be  varied 
without  taking  them  apart,  and  pro- 
vision is  made  for  lifting  the  seats  by 
means  of  a  lever.  The  seats  are  of 
solid  nickel  in  the  best  designs,  to 
minimize  corrosion. 

The  commercial  rating  of  a  safety 
valve  is  based  upon  the  area  exposed 
to  pressure  when  the  valve  is  closed. 

The  number  and  size  of  safety  valves 
FIG.  391.   Consolidated  Pop  Safety  for  a  given  boiler  are  ordinarily  spec- 
ified by  city  or  state  legislation. 

The  logical  method  for  determining  the  size  of  safety  valves  is  to 
make  the  actual  opening  at  discharge  sufficient  to  take  care  of  all 
steam  generated  at  maximum  load.  Most  rules,  however,  are  empir- 
ical and  based  on  the  extent  of  grate  surface,  thus: 

According   to   the   Boiler   Inspection   Department 
Philadelphia,  _     22.5  G 

in  which  P  +  8.6 

A  =  area  of  combined  safety  valves,  inches. 
G  =  Grate  area,  square  feet. 

p  =  boiler  pressure,  pounds  per  square  inch  gauge. 
According  to  the  rule  of   United  States  Supervising  Inspectors   of 
Steam  Vessels,* 


of   the    city   of 
(166) 


A  =  — ,  for  lever  safety  valves. 

/"y 

A  =  — ,  for  pop  safety  valves. 

3 

Other  notations  as  in  (166). 
Button's  rule  is  ("  Steam-Boiler  Construction,")  p.  470  : 


(167) 
(169) 


(169) 


All  notations  as  above. 
*  Superseded  1908  by  the  following: 

A  =  0.2074  ~ 

A  as  above  r 

W  =  weight  of  steam  per  hour,  Ibs. 
P  =  Absolute  steam  pressure. 
See  Power,  Mar.  9,  1908,  p.  480;  Mar.  16,  1909,  p.  520. 


PIPING  AND  PIPE  FITTINGS 


665 


The  Consolidated  Safety  Valve  Company's   circular  gives  the  fol- 
lowing rated  capacity  of  its  nickel-seat  pop  safety  valves: 


Sizes  in  Inches. 

1 

H 

l* 

2 

2* 

3 

3* 

4 

4 

5 

fif 

6 

Boiler  Horse-power  — 
From  

8 

10 

?0 

35 

60 

75 

100 

125 

150 

175 

200 

fl75 

To 

10 

15 

30 

50 

75 

100 

125 

150 

175 

200 

275 

300 

351.  Back-Pressure  and  Atmospheric  Relief  Valves.  —  These  valves 
are  for  the  purpose  of  preventing  excessive  back  pressure  in  exhaust 
pipes.  In  non-condensing  plants  such  valves  are  designated  as  back- 
pressure valves  and  in  condensing  plants  as  atmospheric  relief  valves.  In 
the  former  the  valve  is  usually  adjusted  so  that  a  pressure  of  one  to 


FIG.  392.     Foster  Back-Pressure  Valve. 


FIG.  393.     Davis  Back-Pressure  Valve. 


five  pounds  above  the  atmosphere  is  necessary  to  lift  it  from  its  seat; 
in  the  latter  the  valve  lifts  at  about  atmospheric  pressure.  They  are 
practically  identical  in  construction,  differing  only  in  minor  details. 
A  slight  leakage  in  the  back-pressure  valve  is  of  small  consequence, 
but  in  an  atmospheric  relief  valve  it  may  seriously  affect  the  degree 
of  vacuum  and  throw  unnecessary  work  upon  the  air  pump,  hence  it 
is  customary  to  " water-seal"  the  latter.  Fig.  392  shows  a  section 
through  a  typical  back-pressure  valve.  The  valve  proper  consists  of 
a  single  disk  moving  vertically.  The  valve  stem  is  in  the  form  of  a 
piston  or  dashpot  which  prevents  sudden  closing  or  hammering.  The 
pressure  holding  the  valve  against  its  seat  is  regulated  by  a  spring. 
When  the  back  pressure  becomes  greater  than  atmospheric  plus  that 
added  by  the  spring,  the  valve  raises  from  its  seat  and  relieves  it. 

Fig.  393  shows  a  section  through  a  Davis  back-pressure  valve,  in  which 
the  resisting  pressure  is  varied  by  means  of  a  lever  and  weight. 


666 


STEAM  POWER  PLANT  ENGINEERING 


Fig.  363  shows  the  application  of  a  back-pressure  valve  to  a  typical 
heating  system. 

Fig.  394  shows  a  section  through  a  typical  atmospheric  relief  valve. 
Opening  B  is  connected  to  the  exhaust  pipe  and  opening  A  leads  to 
the  atmosphere.  Under  normal  conditions  of  operation  atmospheric 
pressure  holds  valve  V  against  its  seat.  Water  in  groove  S  "  water- 
seals  "  the  seat  and  prevents  air  from  being  drawn  into  the  condenser. 
In  case  the  pressure  in  pipe  B  becomes  greater  than  atmospheric  it 
lifts  valve  V  from  its  seat  and  is  relieved.  Piston  P  acts  as  a  dash- 
pot  and  prevents  the  valve  from  slamming. 

Fig.  395  shows  a  section  through  an  atmospheric  relief  valve  in 
which  the  weight  of  the  valve  is  counterbalanced  or  even  over- 
balanced by  an  adjustable  weight  and  lever,  thereby  permitting  the 
valve  to  open  at  or  below  atmospheric  pressure,  as  may  be  desired. 


FIG.  394.   Crane  Atmospheric  Relief  Valve.       FIG.  395.   Acton  Atmospheric  Relief  Valve. 


352.  Reducing  Valves.  —  It  is  often  necessary  to  provide  steam  at 
different  pressures  in  the  same  plant,  as  in  the  case  of  a  combined 
power  and  heating  plant.  To  effect  this  result  the  reduction  in  pres- 
sure is  accomplished  by  passing  the  steam  through  a  reducing  valve, 
which  is  but  an  automatically  operated  throttle  valve.  There  are 
many  different  forms,  the  operation  of  all  being  based  upon  the  same 
general  principles. 

In  the  Kieley  valve,  Fig.  396,  the  low-pressure  steam  acts  upon  the 
top  of  flexible  diaphragm  D,  and  the  weighted  lever  L  (which  may  be 
adjusted  to  give  the  desired  reduction  in  pressure)  acts  upon  the  other 
side.  The  movement  of  the  diaphragm  causes  the  balanced  valve  V 


PIPING  AND  PIPE  FITTINGS 


667 


at  the  upper  end  of  the  spindle  to  open  or  close,  as  may  be  necessary 
to  maintain  the  desired  lower  pressure.  Inertia  weights  T  and  C 
prevent  chattering. 


FIG.  396.    Kieley  Reducing 
Valve. 


FIG.  397.    Foster  Pressure 
Regulator. 


Fig.  397  shows  a  section  through  a  class  G  Foster  presrsure  regulator 
or  reducing  valve.  In  operation,  steam  enters  at  A  and  passes 
through  the  main  valve  port  H  to  the  outlet  B.  Steam  at  initial 
pressure  passes  through  port  C  to  chamber  P  and  thence  to  the  top 
of  piston  T  through  port  L,  opening  the  main  valve  U.  Steam  at 
delivery  pressure  passes  through  E  and  raises  the  diaphragm  V 
against  the  pressure  of  spring  R,  allowing  spring  W  to  close  the  aux- 
iliary valve  X.  The  pressure  in  chamber  J  is  then  equalized  by  the 
reduced  pressure  in  ports  G  and  the  under  side  of  piston  X,  and  thus 
allows  spring  Y  to  close  the  main  valve,  which  is  then  held  to  its  seat 
by  the  initial  pressure.  Any  reduction  in  delivery  pressure  is  trans- 
mitted to  diaphragm  V,  and  permits  spring  to  open  auxiliary  valve  X, 
thereby  admitting  steam  to  the  top  of  piston  T,  as  previously 
explained.  The  delivery  pressure  is  adjusted  by  screw  D;  thus 
increasing  the  tension  of  spring  R  increases  the  discharge  pressure 
and  vice  versa.  The  adjustment  once  made,  the  delivery  pressure 


668  STEAM  POWER  PLANT  ENGINEERING 

will  remain  constant,  regardless  of  any  variable  volume  of  discharge 
or  of  the  initial  pressure,  so  long  as  the  latter  is  in  excess  of  the 
delivery  pressure.  W,  Fig.  366,  shows  the  application  of  a  reducing 
valve  to  an  exhaust  steam  heating  system.  Live  steam  is  led  to  the 
valve  through  pipe  A.  It  will  be  noted  that  the  pipe  leading  from 
the  valve  to  the  heating  system  is  much  larger  than  the  high-pressure 
supply  pipe  on  account  of  the  increase  in  volume  of  the  low-pressure 
steam.  Reducing  valves  should  always  be  by-passed  to  permit  of 
repairs  without  shutting  down  the  system.  Care  should  be  taken  in 
not  selecting  too  large  a  reducing  valve,  as  the  valve  lift  is  very  small 
and  the  larger  the  valve  the  less  will  be  the  lift  for  a  given  weight  of 
flow  and  consequently  the  greater  the  wire  drawing  and  erosion  of 
the  valve  seat. 

353.  Foot  Valves.  —  Whenever  a  long  column  of  water  is  to  be 
moved  in  either  suction  or  delivery  pipe  it  is  customary  to  place  a 
check  valve  near  the  lower  end  of  the  column  to  prevent  the  water 
from  backing  up  when  the  pump  reverses  or  shuts  down.  The  check 
valve  placed  at  the  end  of  the  suction  pipe  is  called  a  foot  valve. 
Any  check  valve  may  be  used  as  a  foot  valve,  though  practice  limits 
the  choice  to  the  disk  or  flap  type  as  illustrated  in  Fig.  398.  To  pre- 
vent rubbish  from  destroying  the  action,  a  strainer  or  screen  is  gener- 


(B) 

FIG.  398.     Types  of  Foot  Valves. 

ally  incorporated  with  the  body  of  the  valve.  A,  Fig.  398,  illustrates  a 
single-flap,  B  a  multi-flap  and  C  a  disk  valve  composed  of  a  nest  of 
small  rubber  valves.  The  single-flap  are  usually  made  in  sizes  J  to 
6  inches,  the  multi-flap  7  to  16  inches,  and  the  disk  valve  in  all  com- 
mercial sizes  from  f  to  36  inches.  For  large  sizes,  16  to  36  inches,  the 
multi-disk  valve  is  given  preference,  since  a  number  of  the  disks  may 
be  disabled  without  destroying  its  operation. 

The  Use  and  Abuse  of  Globe  Valves:  Power  &  Engr.,  Jan.,  1909,  p.  10. 
Gate  Valves  in  Steam  Pipe  Lines:  Power  &  Engr.,  Feb.  16,  1909,  p.  320. 
Types  of  Check  Valves  and  their  Operation:  Power  &  Engr.,  July  6,  1909,  p.  11. 


CHAPTER  XYI. 

LUBRICANTS  AND  LUBRICATION. 

354.  General.  —  The  losses  due  to  the  friction  of  the  working  parts 
of  machinery  include  considerably  more  than  the  mere  loss  of  power, 
namely,   the   depreciation   resulting  from   wear   of   bearings,    guides, 
and  other  rubbing  surfaces,  and  the  expense   arising  from  accidents 
traceable  to  excessive  friction.     The  power  absorbed  in  overcoming 
friction  varies  with  the  type  of  plant  and  the  character  of  machinery 
and  is  seldom  less  than  5  per  cent  and  often  greater  than  30  per  cent 
of  the  total  power  developed.     In  large  central  stations  these  losses 
approximate  8  per  cent  and  in  weaving  and  spinning  mills  will  average 
as  high  as  25  per  cent.     (Trans.  A.S.M.E.,  6-465.)     These  figures  refer 
to  properly  lubricated  plants  operating  under  normal  conditions.     The 
proper  selection  of  lubricant  is  therefore  a  very  important  problem, 
since,  besides  the  cost  of  the  lubricant  itself,  the  loss  in  power  and  in 
wear  and  tear  to  machinery  is  no  small  item.     A  change  of  lubricant 
may  frequently  result  in  marked  increase  in  economy  of  operation. 
The  lubricants  most  commonly  met  with  in  power  plant  practice  are 
conveniently  classified  as  oils,  greases,  and  solids,  and  are  of  animal, 
mineral,  or  vegetable  origin. 

Reference  Books:  Archbutt  and  Deeley,  Lubrication  and  Lubricants;  Redwood 
Lubricants;  W.  M.  Davis,  Friction  and  Lubrication;  Gill,  Oil  Analysis;  Robinson, 
Gas  and  Petroleum  Engines;  Thurston,  Friction  and  Lost  Work;  Gill,  Engine  Room 
Chemistry. 

355.  Vegetable  Oils.  —  Except  for  certain  special  purposes  and  for 
compounding  with  mineral  oils  these  possess  lubricating  properties  of 
little  practical  value,  since  they  decompose  at  comparatively  low  tem- 
peratures and  have  a  tendency  to  become  thick  and  gummy.     The  vege- 
table oils  sometimes  employed  are  linseed,  cottonseed,  rape,  and  castor. 

Vegetable  Oils:  Power,  May,  1906,  p.  300;  Archbutt  and  Deeley,  Lubrication 
and  Lubricants,  p.  232;  W.  M.  Davis,  Friction  and  Lubrication,  p.  28;  Gill,  Oil 
Analysis. 

356.  Animal   Fats.  —  Many   animal   fats   have   greater   lubricating 
power  than  pure  mineral  oils  of  corresponding  viscosity  but  are  objec- 
tionable  on  account  of  their  unstable  chemical  composition.      They 
decompose  easily,  especially  in  the  presence  of  heat,  and  set  free  acids 


L 


670 


STEAM  POWER  PLANT  ENGINEERING 


which  attack  metals.  They  are  seldom  used  in  the  pure  state  and 
are  usually  compounded  with  mineral  oils.  The  animal  products 
used  in  this  connection  are  tallow,  neat's-foot  oil,  lard,  sperm,  wool 
grease,  and  fish  oil,  the  first  named  being  the  most  important.  In 
cylinder  lubrication,  especially  in  the  presence  of  moisture,  the  addi- 
tion of  2  to  5  per  cent  of  acidless  tallow  seems  to  make  the  oil  adhere 
better  to  the  metal  surfaces  and  increases  the  lubricating  effect,  while 
the  proportion  is  so  small  that  ill  effects  from  corrosion  or  gumming 
are  scarcely  perceptible. 

Animal  Fats:  Archbutt  and  Deeley,  Lubrication  and  Lubricants,  p.  323;  Gill, 
Oil  Analysis,  p.  44;  Wright,  Analysis  of  Oils,  p.  193;  Andes,  Animal  Fats. 

357.  Mineral  Oils.  —  These  are  all  products  of  crude  petroleum  and 
form  by  far  the  greater  part  of  all  lubricants.  They  present  a  wider 
range  of  lubricating  properties  than  those  derived  from  animal  or 
vegetable  sources,  the  thinnest  being  more  fluid  than  sperm  and  the 
thickest  more  viscous  than  fats  and  tallows.  They  are  not  easily 
oxidized,  do  not  decompose,  become  rancid,  or  contain  acids. 

Crude  American  petroleum  of  specific  gravity  0.802  may  yield 
the  following  commercial  products.  ("  Gas  and  Petroleum  Engines," 
W.  Robinson.) 


Average 
Percentage. 

Boiling 
Specific          Point, 
Gravity.        Degrees 
F. 

Light  Oils. 

(  Cymogene       

traces 

0  590            32 

Petroleum  ether   

•]  Rhigolene  

0  1 

625-  631        64 

f  Gasoline  
(  C.  Naphtha  

1-1.5 
10 

.635-.  658      85-155 
680-  700    140-212 

Petroleum  spirit 

<  B    Naphtha 

2-2  5 

717-  72      175-245 

Burning  oils,  kerosene. 
Fuel  oils 

(  A.  Naphtha  (benz.)  .... 

(Water  white  
\  Ordinary  kerosene  

/For  making  oil  gas;  fuel 

2-2.5 

12-20 
40-55 

.742-.  745    212-265 

.780-.  785    300-570 
.800-.  810    300-680 
and  up- 
wards 
0  85 

Heavy  oils 

(  Lubricating  oils  
<  Paraffin  wax  

17.5 
2 

.885-.  920    480  and 
908  at  60  upwards 

'  Residium 

5-10 

deg    F 

Mineral  lubrication  oils  may  be  classified  as 

(1)  Distilled  oils,  which  are  produced  by  distillation  from  crude 
petroleum  and  made  pale,  amber  colored,  and  transparent  by  treat- 
ment with  acid  and  alkali. 


LUBRICANTS  AND  LUBRICATION  671 

(2)  Natural  oils,  which  are  prepared  from  crude  petroleum,  from 
which  grit,  suspended  and  tarry  impurities  have  been  removed.     They 
are  dark  and  opaque  and  are  rich  in  lubricating  properties. 

(3)  Reduced  oils,  or  heavy  natural  oils,  from  which  the  lighter  hydro- 
carbons have  been  evaporated  and  from  which  the  tarry  residue  has 
been  removed  by  filtration. 

Mineral  Lubricants:  Engr.  U.S.,  July  1, 1904,  pp.  466,  Vol.  44  (1907),  pp.  241,  369, 
542,  585;  National  Engr.,  Jan.,  1905,  p.  19;  Eng.  Mag.,  June,  1904,  p.  455;  Power, 
March,  1906,  p.  146. 

358.  Solid    Lubricants.  —  Dry    graphite,    soapstone,  and    mica   are 
sometimes  used  as  lubricants,  though  they  are  usually  mixed  with 
grease   or   oils.     They   cannot    easily   be   squeezed   or   scraped   from 
between  the  surfaces,  and  are  consequently  suitable  where  very  great 
weights  have  to  be  carried  on  small  areas  and  when  the  speed  of  rub- 
bing is  not  high.     The  coefficient  of  friction  of  such  lubricants  is  high, 
and  when  economy  of  power  is  essential  better  results  may  be  secured 
by  the  use  of  liberally  proportioned  rubbing  surfaces  and  liquid  lubri- 
cants.     Under  certain  conditions  of  pressure  and  speed  these  lubri- 
cants will  sustain,  without  injury  to  the  surfaces,  pressures  under  which 
no  liquid  would  work. 

Graphite:  Trans.  A.S.M.E.,  13-374;  Engng.,  Aug.  16,  1907;  Sci.  Am.,  May  11, 
1907;  National  Engr.,  Jan.,  1904;  Am.  Mach.,  Dec.,  1907,  pp.  784,  934;  Horseless 
Age,  Jan.,  1904,  June  11,  1902,  p.  712;  Power,  Dec.,  1906,  p.  758. 

359.  Greases.  —  Under   this    name   may   be   included   the    various 
compounds  which  consist  of  oils  and  fats  thickened  with  sufficient 
soap  to  form,  at  ordinary  temperatures,  a  more  or  less  solid  grease. 
Those  usually   employed   are  lime,  soda,   or  lead  soaps,  made   with 
various  fats  and  oils.     "  Engine  "  greases  are  thickened  with  a  soap 
made  from  tallow  or  lard  oil  and  caustic  soda,   and  often  contain 
neat's-foot  oil,  beeswax,  and  the  like.     For  exceptionally  heavy  pres- 
sures, graphite,   soapstone,   and  mica   are   sometimes    added   to   the 
grease. 

Greases:  Jour.  Eng.  Soc.  West.  Penn.,  March,  1904,  p.  112;  Railroad  Gazette, 
July  8,  1904,  p.  131;  St.  Ry.  Jour.,  July,  1905,  p.  95;  see  also  text-books  given  in 
references  at  beginning  of  chapter. 

360.  Qualifications  of  Good  Lubricants.  —  A  good  lubricant  should 
possess  the  following  qualities: 

(1)  Sufficient  "  body  "  to  prevent  the  surfaces  from  coming  into 
contact  under  conditions  of  maximum  pressure. 

(2)  Capacity  for  absorbing  and  carrying  away  heat. 


672  STEAM  POWER  PLANT  ENGINEERING 

(3)  Low  coefficient  of  friction. 

(4)  Maximum  fluidity  consistent  with  the  "  body  "  required. 

(5)  Freedom  from  any  tendency  to  oxidize  or  gum. 

(6)  A  high  "  flash  point  "  or  temperature  of  vaporization  and  a  low 
congealing  or  "  freezing  point." 

(7)  Freedom    from    corrosive    acids    of    either    metallic    or    animal 
origin. 

Lubricating  oils  are  identified  by  certain  tests  which  are  used  by 
refiners  in  grading  and  classifying  the  oils  and  by  consumers  in  buy- 
ing them.  These  tests  usually  cover  the  following: 

(1)  Identification  of  the  oil,  whether  a  simple  mineral,  animal  or 
vegetable  oil  or  a  mixture. 

(2)  Density  or  gravity. 
(3.)  Viscosity. 

(4)  Flash  point. 

(5)  Burning  point,  fire  test. 

(6)  Acidity. 

(7)  Coefficient  of  friction. 

(8)  Cold  test. 

361.  Identification  of  Oil.  —  The  chemical  analysis  of  oils  lies  in 
the  province  of  the  chemist,  but  some  of  the  characteristics  may  be 
readily  determined  by  a  few  simple  tests.  To  detect  admixtures  of 
fatty  oils  in  mineral  oil  a  small  quantity  is  heated  in  a  test  tube  for 
15  minutes  with  small  pieces  of  either  metallic  sodium  or  caustic 
potash.  If  fatty  oil  is  present,  saponification  takes  place  and  the 
soap  formed  will  rise  to  the  top  as  a  semi-solid  mass  and  the  amount 
may  be  estimated.  Tarry  matter  may  be  detected  by  dissolving  a 
small  quantity  of  oil  in  from  10  to  20  times  its  bulk  of  gasoline;  the 
tar  and  other  insoluble  matter  will  separate  and  collect  at  the  bottom. 

Oil  Testing  and  Specifications:  Power,  May,  1904,  p.  302,  Vol.  24  (1904),  pp.  139, 
240,302,  526,  Vol.  26  (1906),  pp.  145,  222,  300,  331,  407;  Am.  Mach.,  April  11,  1907, 
p.  525;  Engr.  U.S.,  Oct.  15,  1904,  p.  724,  Oct.  2,  1905,  p.  657;  Marine  Engng.,  June, 
1903,  p.  303;  Chem.  Engr.  Nov.,  1905,  p.  10,  Dec.,  1905,  p.  87,  Jan.,  1906,  p.  141; 
Am.  Gas  Light  Jour.,  Jan.  23,  1905;  U.S.  Cons.  Repts.,  June,  1905;  Sci.  Am.  Sup., 
Jan.  14,  1905. 

363.  Gravity.  —  The  density  or  specific  gravity  is  conveniently 
determined  by  means  of  a  hydrometer,  which,  in  the  oil  trade,  is 
graduated  according  to  the  Baume*  scale.  The  relationship  between 
specific  gravity  and  degrees  Baume"  at  a  temperature  of  60  degrees  F. 
may  be  expressed: 

140 

Specific  gravity  = ; — —  • 

130  +  degrees  BaumS 


LUBRICANTS  AND  LUBRICATION 


673 


Table  88  gives  the  specific  gravity  and  gravity  Baum<§  of  a  number 
of  lubricating  oils. 

Gravity:  Power,  March,  1904,  p.  139;  Robinson,  Gas  and  Petroleum  Engines, 
p.  474;  Archbutt  and  Deeley,  Lubrication  and  Lubricants,  pp.  172-185;  W.  M.  Davis, 
Friction  and  Lubrication,  p.  34. 

TABLE  88. 

SPECIFIC  GRAVITY  AND  GRAVITY  BAUME  OF  A  NUMBER  OF  LUBRICANTS. 


Specific  Grav- 
ity. 

Gravity 
Baumg. 

Flash  Test, 
Degrees  F. 

Water 

1  000 

10 

Cylinder  oil  

.9090 

24.5 

575 

Cylinder  oil  

.8974 

26 

540 

Heavy  engine  oil                              

.9032 

25.5 

411 

Medium  engine  oil                 .    .    .       

.9090 

24 

382 

Light  engine  oil            

.8917 

27 

342 

Castor  machine  oil       

.8919 

27 

324 

.9175 

23 

505 

Sperm  oil         

.8815 

29 

478 

Tallow  oil         

.9080 

24.5 

540 

Cottonseed  oil  

.9210 

22 

518 

Linseed  oil     

.9299 

19 

505 

Castor  oil  (pure) 

9639 

15 

Palm  oil 

9046 

25 

405 

Rape  seed  oil 

9155 

23 

Spindle  oil  

.8588 

33 

312 

363.  Viscosity.  —  Viscosity  may  be  defined  as  the  degree  of  fluidity 
or  internal  friction  of  an  oil.     It  is  sometimes  called  the  "  body."     It 
is  determined  by  a  viscosimeter.     There  are  a  number  of  different 
instruments  for  this  purpose  but  no  recognized  standard  instrument  or 
method,  so  that  "  viscosity  "  conveys  no  meaning  unless  the  name  of 
the  instrument,  the  temperature,  and  the  amount  of  oil  tested  are 
given.     Nearly  all  instruments  are  of  the  orifice  type;    that  is,  the 
viscosity  of  an  oil  is  taken  as  the  number  of  seconds  required  for  a 
given    amount,    usually   50   cubic    centimeters,    to   flow    through   an 
orifice  at  a  given  temperature.     By  "  specific  viscosity  "  is  meant  the 
ratio  of  the  time  required  for  the  oil  to  run  out  to  that  of  an  equal 
quantity  of  water  at  60  degrees  F.     The  viscosity  of  engine  oils  is 
usually  taken  at  70  degrees  F.  and  of  cylinder  oils  at  212  degrees  F. 

Viscosity:  Trans.  A.S.M.E.,  9-369;  Engr.,  Lond.,  Sept.  7,  1906,  p.  344,  June  12, 
1900,  p.  633;  Eng.  Mag.,  June,  1907,  p.  455;  Machinery,  May,  1903,  p.  484;  Power, 
May,  1904,  p.  303,  May,  1907,  p.  293,  March,  1906,  p.  146. 

364.  Flash  Point.  —  The  flash  point  is   determined  by  heating  a 
sample  of  oil  in  an  open  or  closed  cup  at  the  rate  of  15  degrees  F.  per 
minute  until  a  spark  will  ignite  the  vapor.     The  temperature  at  which 


674  STEAM  POWER  PLANT  ENGINEERING 

this  occurs  is  the  flash  point.  So  much  depends  upon  the  extent  of  oil 
surface  exposed,  size  of  spark,  distance  spark  is  held  from  the  oil  at  the 
time  of  ignition,  and  the  dimensions  of  the  cup,  that  there  may  be  con- 
siderable variation  in  the  flash  point  as  obtained  by  different  experi- 
menters. 

Flash  Test:  Power,  April,  1906,  p.  222;  Robinson,  Gas  and  Petroleum  Engines, 
pp.  482-488;  Archbutt  and  Deeley,  Lubrication  and  Lubricants,  pp.  187-191;  W.  M. 
Davis,  Friction  and  Lubrication,  p.  34;  Gill,  Oil  Analysis,  p.  36. 

365.  Burning  Point,  or  Fire  Test.  —  By  continuing  the  application 
of   heat  and  noting  the  temperature  at  which   the  oil  takes  fire  and 
continues   to   burn,  the  burning  point  is  obtained.     The  higher  the 
temperature  under  which  the  oil  must  work  the  higher  the  fire   test 
required,  so  that  it  will  not  decompose  or  volatilize.     Too  high  a  fire 
test  gives  an  oil  that  does  not  atomize  readily  enough  to  reach  all  parts 
of  the  cylinder. 

Consult  references  under  "  Flash  Test." 

366.  Acidity.  —  The  presence  of  free  acid  is  determined  by  shaking 
up  equal  quantities  of   oil  and  water  and  testing  with  litmus  paper. 
Another  simple  test  is  as  follows:    A  small  quantity  of  oil  is  placed  in 
a  test  tube  with  a  little  cupric  oxide  (Cu2O)  and  subjected  to  a  gentle 
heat  for  three  or  four  hours.     The  reaction  with  the  copper  turns  the 
solution  green  if  fatty  acid  is  present  and  blue  if  vegetable  acid  is 
present. 

Acidity:  Power,  April,  1906,  p.  222;  Archbutt  and  Deeley,  Lubrication  and 
Lubricants,  pp.  215-218;  Gill,  Oil  Analysis,  p.  74. 

367.  Cold  Test.  —  The  "  cold  test  "  is  the  temperature  at  which  the 
oil  will  just  flow.     The  sample  is  solidified  by  means  of  a  freezing 
mixture  and  the  temperature  noted  when  it  softens  sufficiently  to  flow. 

Cold  Test:  Robinson,  Gas  and  Petroleum  Engines,  p.  481 ;  Archbutt  and  Deeley, 
Lubrication  and  Lubricants,  pp.  195,  200-6;  W.  M.  Davis,  Friction  and  Lubrication, 
p.  28;  Gill,  Oil  Analysis,  p.  34;  Redwood,  Lubricants,  p.  3;  Power,  March,  1906,  p.  146. 

368.  Friction    Test.  —  The    coefficient    of    friction    as    determined 
from  friction-testing  machines  is  useful  in  obtaining  a  comparison  of 
oils  under  the  test  conditions,  but  gives  little  information  concerning 
the  action  of  the  oil  under  the  widely  different  conditions  found  in 
actual  practice.     Table  89  gives  the  physical  properties  of  a  number 
of  lubricating  oils,  with  their  particular  zone  of  application. 

Friction  and  Lubrication:  Trans.  A.S.M.E.,  1-74,  6-136;  Am.  Mach.,  July  21, 
1904,  p.  956,  Jan.  23,  1902,  p.  113,  Sept.  10,  1903,  p.  1303;  Am.  Elecn.,  Nov.,  1905, 
p.  557;  Engr.,  Lond.,  June  19,  1903,  p.  631;  Power,  Dec.,  1905,  p.  748;  National 
Engineer,  Jan.,  1905,  p.  19;  Mech.  Engr.,  Nov.  30,  1907;  Pro.  Inst.  Civ.  Engr.,  1901, 
p.  146;  Machinery,  Aug.,  1903,  p.  631 ;  Proc.  A.S.M.E.,  Nov.,  1909,  p.  1099. 


LUBRICANTS  AND  LUBRICATION  675 

369.  Atmospheric  Surface  Lubrication.  —  In  a  general  sense  all 
journals,  slides,  and  "  atmospheric  "  surfaces  should  be  lubricated  with 
straight  mineral  oils  (as  free  from  paraffin  as  possible),  except  when 
in  contact  with  considerable  water,  in  which  case  it  is  advisable  to  add 
20  to  30  per  cent  of  lard  oil.  Vegetable  oils,  paraffin  oils,  and  animal 
oils  (except  lard  oil  as  above  stated)  are  not  recommended  for  general 
engine  and  dynamo  service.  The  test  requirements  of  a  number  of 
classes  of  lubricants  are  outlined  in  Table  89  and  represent  current 
practice.  Bearings,  guides,  and  all  external  rubbing  surfaces  may  be 
lubricated  in  a  number  of  ways.  (1)  They  may  be  given  an  inter- 
mittent application  of  oil,  as,  for  example,  with  an  oil  can;  (2)  they  may 
be  equipped  with  oil  cups  with  restricted  rates  of  feed;  and  (3)  they 
may  be  flooded  with  oil.  The  relative  lubricating  values  of  the  systems 
have  been  estimated  approximately  as  follows  (Power,  December,  1905, 
p.  750) : 


Coefficient  of  Fric- 
tion. 

Comparative 
Value. 

Intermittent 

0  01  and  greater 

72  and  less 

Restricted  feed 

0  01  to  0  012 

79  to  86 

Flooded  bearing 

0  00109 

100 

370.  Intermittent  Feed.  —  Intermittent   applications   are  ordinarily 
limited  to  small  journals,  pins,  and  guides  which  are  subject  to  light 
pressures  and  which  do  not  easily  permit  of  oil  or  grease  cups,  as,  for 
example,  parts  of  the  valve  gear  of  a  Corliss  engine,  governors  and 
link  work.     On  account  of  the  labor  attached  and  the  frequent  doubt 
about  the  oil  reaching  the  wearing  surfaces  this  method  of  lubrication 
is  limited  as  much  as  possible  even  in  the  smallest  plants. 

371.  Restricted  Feed.  —  In  the  average  power  plant  the  major  part 
of  the  lubrication  is  effected  by  means  of  oil  cups  which  are  filled  at 
intervals  by  hand  or  by  mechanical  means,  the  oil  being  fed  from  the 
cup  by  drops,  according  to  the  requirements. 

372.  Oil  Bath.  —  In  large  power  plants  the  principal  journals  and 
wearing  parts  are  supplied  with  a  continuous  flow  of  oil  which  com- 
pletely "  floods  "  the  rubbing  surfaces.     The  oil  is  forced  to  the  various 
parts  either  by  gravity  from  an  elevated  tank  or  by  pressure  from  a 
pump.     After  the  oil  leaves  the  bearings  it  flows  into  collecting  pans, 
thence  into  a  receiving  and  filtering  tank,  and  finally  is  pumped  back 
into  an  elevated  reservoir  and  used  over  and  over  again.     The  little 
lost  by  leakage  and  depreciation  is  replenished  by  the  addition  of  new 
oil  to  the  system. 


676 


STEAM  POWER  PLANT  ENGINEERING 


TABLE   89. 
PHYSICAL    CHARACTERISTICS    OF    A    NUMBER    OF    LUBRICANTS. 

(Power,  December,  1905,  p.  750.) 


*t 

I  8 

1     CO 

1   ? 

^ 

Kind  of  Oil. 

Use  and  Adaptation. 

•~  S, 

H  fe 

JO    & 

8  °  1 

0° 

s* 

&* 

SQ 

!«& 

High-pressure  cylinder 

For  steam  cylinders  using  dry 

25 

600 

645 

175 

oil. 

steam  at  pressures  from  110 

to 

30 

to 

to 

to 

to  210  pounds. 

24.5 

610 

660 

205 

General  cylinder  oil  .  . 

For  steam  cylinders  using  dry 

steam  at  75  to  100  pounds. 
For  air  compressor  cylinders 

26 
to 

30 

550 
to 

600 
to 

180 
to 

when   made  from   steam-re- 
fined mineral  stock  and  when 

25.5 

585 

630 

190 

viscosity  is  200. 

Wet  cylinder  oil. 

For  use  where  the  steam  is  moist, 

25.8 

560 

600 

150 

(Remark  1.) 

especially  in  compound  and 

to 

30 

to 

to 

to 

triple  expansion  engines. 

25.3 

585 

630 

185 

Gas  engine  cylinder  oil. 

For  gas  engine  cylinders.     Neu- 

(Remark 2.) 

tral  mineral  oil  compounded 
with  an  insoluble  soap  to  give 

26.5 

30 

320 

350 

300 

body. 

Automobile  gas  engine 
oil.     (Remark  3.) 

For  automobile  gas  engines  and 
similar  work. 

29.5 

30 

430 

485 

195 

Heavy     engine     and 

For  heavy  slides  and  bearings, 

30.5 

440 

170 

machinery  oils. 

shafting,  and  horizontal  sur- 

to 

30 

400 

to 

to 

faces. 

29.5 

450 

195 

General    engine    and 

For    high-speed   dynamos   and 

30.8 

400 

450 

175 

machine  oils. 

.  machines. 

to 

30 

to 

to 

to 

30 

420 

470 

190 

Fine  and  light  machine 

For    fine  work,   from   printing 

oils. 

presses  to   sewing   machines 

00       C 

i  in 

and  typewriter  oils.     With  a 
cold  test  of  25°  to  28°  and  a 

to 

30 

400 

440 

to 

viscosity  of  140°  this  makes 

30.2 

160 

an  excellent  spindle  oil. 

Cutting  and  heat  dis- 

For cutting  tools,  screw  cutting 

27 

410 

475 

210 

sipating  oils. 

and  similar  work. 

to 

30 

to 

to 

to 

(Remark  4.) 

23 

420 

480 

175 

For  ice  machinery 

30.2 

0 

200 

225 

165 

Wet  service  and  marine 

For  marine  service,  or  where  a 

oils.     (Remark  4.) 

great  deal  of  moisture  must 

28 

30 

430 

475 

230 

be  handled. 

Greases        

They  are  used  in  special  work 

and  for  heavy  pressures  mov- 

ing at  slow  velocities. 

Remark  1.  —  May  contain  not  over  2  to  6  per  cent  of  refined  acidless  tallow  oil  in  the  high- 
pressure  oils  and  not  over  6  to  12  per  cent  in  the  low-pressure  oils. 

Remark  2.  —  The  reason  for  using  an  insoluble  soap  such  as  oleate  of  aluminum  is  that  it 
is  impossible  to  decompose  the  soap  with  a  high  heat ;  the  soap,  although  not  a  lubricant,  is  a 
vehicle  for  carrying  some  oil. 

Remark  3.  —  Owing  to  a  lack  of  body,  this  oil  will  not  interfere  with  the  sparking  by  depos- 
iting carbon  on  the  platinum  point. 

Remark  4.  —  May  contain  30  to  40  per  cent  of  pure  strained  lard  oil. 


LUBRICANTS  AND  LUBRICATION 


677 


373.   Oil  Cups.  —  Fig.  399  illustrates  the  application  of  sight-feed  oil 
cups  to  the  crosshead  and  slides  of  a  reciprocating  engine.     The  oil  is 

fed  into  the  cups  by  hand  and 
gravitates  to  the  rubbing  surfaces, 
the  rate  of  flow  being  regulated  by 
a  needle  valve.  Cups  A  and  B 
feed  directly  to  the  crosshead 
guides,  but  the  oil  from  cup  D 
flows  to  the  bottom  orifice  0, 
from  which  it  is  wiped  by  a  metal- 
lic wick  S  and  carried  by  gravity 
to  the  wrist  pin. 

374.  Telescopic  Oiler. — I  ig.  400 
shows  the  application  of  a  tele- 
scopic oiler  to  a  crosshead  and 
guides.  0  and  C  are  sight-feed 
oil  cups,  the  former  feeding  directly 


FIG.  399. 


Oil  Cup  Lubrication,  Hand 
Filled. 


to  the  top  guide  through  the  tube 
S.     The    oil     from    C    flows    by 

gravity  through  the  swing  joint  into  the  telescopic  tubes  P,  R  and 
thence  to  the  pin  through   the   lower  swing   joint  as   indicated.     As 


FIG.  400.    Nugent  'a  Telescopic  Oiler. 

the  crosshead  moves  back  and  forth,  the  pipe  P  slides  into  and  out  of 
pipe  R,  the  oil  being  thus  conducted  directly  to  the  pin  without  wasting. 


678 


STEAM  POWER  PLANT  ENGINEERING 


A  device  of  this  type  installed  on  a  high-speed  automatic  engine  at  the 
Armour  Institute  of  Technology  has  been  in  operation  for  three  years 
without  cost  for  repair  or  renewal. 

375.   Ring  Oiler.  —  Small  high-speed  engines  are  often  oiled  by  the 
oil-ring  system,  as  illustrated  in  Fig.  401.     The  shaft  is  encircled  by 


FIG.  401.    Oil  Ring  Lubrication. 


FIG.  402.    Centrifugal  Oiler. 


several  loose  rings  which  dip  into  a  bath  of  oil  in  the  base  of  the 
pedestal  or  frame  and,  rolling  on  the  shaft  as  it  turns,  carry  oil  to 
the  top  of  the  shaft  where  it  spreads  to  the  bearings.  In  some  cases 
the  rings  are  replaced  by  loops  of  chain. 

Ring  Lubrication:   Zeit.  d.  Ver.  Deutscher  Ing.,  Aug.  10,  1907. 


FIG.  403.     Pendulum  Oiler. 


376.   Centrifugal   Oiler.  —  Fig.  402  illustrates  the  application   of   a 
centrifugal  oiler  to  a  side-crank  engine.     The  oil  supply  is  regulated  by 


LUBRICANTS  AND  LUBRICATION 


679 


the  sight-feed  cup  C  and  flows  by  gravity  to  the  pipe  P  in  line  with 
the  center  of  the  crank  shaft.  Centrifugal  force  throws  the  oil  out- 
ward through  pipe  B  to  the  center  of  the  pin  D,  which  is  drilled  longi- 
tudinally and  radially  so  as  to  distribute  the  oil  upon  the  bearing 
surface. 

377.  Pendulum  Oiler.  —  Fig.  403  illustrates  the  application  of  a 
pendulum  oiler  to  the  crank  pin  of  a  center-crank  engine.  Oil  cups 
and  pendulum  P  are  fastened  to  the  crank  shaft  S  by  trunnion  T '. 
The  pendulum  holds  the  cup  vertical,  since  the  friction  of  the  trunnion 
is  not  sufficient  to  revolve  it.  Oil  flows  along  the  center  of  the  crank 
shaft  under  the  head  of  oil  in  cups  0  and  is  thrown  outward  to  bear- 
ing B  by  centrifugal  force. 


FLEXIBLE 

FILLING 

PIPE 


BASEMENT     FLOOR  LINE 


FIG.  404.    Simple  Gravity  Feed  System. 

378.   "  Splash"  Oiling.  —  In  some  high-speed  engines  the  crank,  con- 
necting rod,  and  crossheads   are  inclosed  by  a  casing,  the  bottom  of 


680  STEAM  POWER  PLANT  ENGINEERING 

which  is  filled  with  oil  to  such  a  depth  that  at  each  revolution  of  the 
crank,  the  end  of  the  connecting  rod  is  partly  submerged.  The  result. 
is  that  the  oil  is  splashed  into  every  part  of  the  chamber,  and  the 
crank  pin,  crosshead  pin,  and  crosshead  slides  practically  run  in  an  oil 
bath. 

379.  Gravity  Oil  Feed.  —  Fig.  404  illustrates  a  simple  gravity  oil  feed 
system.     The  oil  to  the  engine  is  supplied  from  the  oil  tank  by  pipe  D 
under  pressure  corresponding  to  the  height  of  the  tank  above  the  oil 
cups.     After  performing  its  function  the  oil  gravitates  to  the  filter  and 
from  the  latter  to  the  oil  reservoir,  from  which  it  is  pumped  back  to 
the  supply  tank,  the  overflow  being  returned  to  the  reservoir  through 
pipe  N.     Operation  is  interrupted  only  when  new  oil  is  to  be  added  to 
the  system  from  the  barrel  through  the  flexible  filling  pipe.     In  case 
the  oil  tank  is  put  out  of  commission,  or  the  supply  pipe  becomes 
clogged,  full  pump  pressure  may  be  used  by  closing  valves  R  and  S 
and  opening  valve  E.     The  make-up  oil  is  small  in  amount  compared 
to  the  quantity  circulated.     The  reclaiming  and  purifying  of  the  oil 
are  essential  if  the  bearings  are  to  be  flooded,  otherwise  the  cost  of  oil 
would  be  prohibitive.     At  the  power  house  of  the  South  Side  Elevated 
Railway  the  daily  circulation   (24  hours)    of  engine  oil  is   approxi- 
mately 1500  gallons.     The  make-up  oil  amounts  to  eight  gallons. 

An  objection  sometimes  made  to  the  above  system  is  that  the  vary- 
ing heights  of  oil  in  the  supply  tank  may  cause  considerable  variation 
in  pressure  at  the  oil  cups,  causing  them  to  feed  faster  when  the  tank 
is  full  and  slower  when  the  tank  is  nearly  empty.  This  applies  only 
to  installations  where  the  supply  tank  is  filled  intermittently. 

380.  Low-Pressure  Gravity  Feed.  —  Fig.  405  shows  the  application 
of  a  low-pressure  oiling  system  in  which  the  level  in  the  sight  feeds  is 
kept  constant.     A  is  the  main  supply  tank,  B1  and  B2  the  upper  and 
lower  gauges  indicating  the  oil  level,  C  the  supply  pipe  running  to  the 
engines,  and  D  a  small  standpipe  closed  at  one  end  and  vented  near 
the  top.   The  reservoir  is  supplied  with  oil  by  the  valve  marked  "  inlet." 
When  the  tank  is  filled  the  oil  rises  in  the  standpipe  D  a  correspond- 
ing height.     The  inlet  valve  is  then  closed  and  the  oil  in  the  standpipe 
feeds  down  to  the  level  of  the  sight  feeds  or  to  a  point  where  the  air 
will  enter  the  bottom  of  the  tank.     This  will  be  the  constant  oil  level, 
since  oil  flows  from  the  tank  only  in  proportion  to  the  amount  of  air 
admitted.     A  head  of  6  inches  has  been  found  to  give  the  best  results. 
(Engineer,  U.  S.,  March  16,  1903,  p.  243.) 

381.  Compressed-Air  Feed.  —  Fig.  406  shows  diagrammatically  the 
arrangement  of  the  oiling  system  at  the  First  National  Bank  Building, 
Chicago.     The  storage   tank  containing  the  supply  of  engine  oil   is 


LUBRICANTS  AND  LUBRICATION 


681 


FIG.  405.    Low-Pressure  Gravity  Feed,  Constant  Head. 


WASTE 


i  t 


OIL  STORAGE 


7r///^ 

FIG.  406.     Oiling  System  at  the  Power  Plant  of  the  First  National  Bank  Building,  Chicago 


682  STEAM  POWER  PLANT  ENGINEERING 

under  air  pressure  at  all  times  except  during  the  short  periods  when  it 
is  being  filled  with  oil  from  the  filter.  The  air  pressure  on  the  surface 
of  the  oil  forces  it  to  a  manifold  on  the  engine  from  which  it  is  dis- 
tributed to  the  various  oil  cups.  The  oil  flows  from  the  different 
bearings  to  the  returns  tank  located  at  the  base  of  the  engines.  When 
the  tank  is  filled  air  pressure  is  admitted  and  the  oil  forced  to  the 
settling  tank,  which  has  a  capacity  of  about  400  gallons  and  is  located 
near  the  ceiling.  The  oil  is  allowed  to  settle  and  the  entrained  water 
and  foreign  material  are  drained  to  waste.  The  oil  gravitates  from  this 
tank  to  a  series  of  Turner, oil  filters.  When  a  new  supply  of  oil  is 
needed,  valves  A  and  B  are  closed  and  vent  valve  C  opened,  cutting 
off  the  supply  of  air  and  reducing  the  pressure  to  atmospheric.  Valve  D 
is  then  opened  and  oil  flows  from  the  filters  to  the  storage  tank. 

Lubricating  Systems.  —  Lubrication  of  Line  and  Counter  Shafting:  Trans.  A.S.M.E., 
10-810.  Gravity  Oil  Systems:  Power,  Nov.,  1902,  p.  23,  July,  1906,  p.  409,  June, 
1903,  p.  305.  Oiling  Systems  for  Electric  Engines:  Elec.  World,  July  7,  1906,  p.  26. 
Oiling  System  for  Power  Plants:  Engr.  U.S.,  March  16,  1903,  p.  243,  April  15,  1904, 
p.  278;  National  Engr.,  Feb.,  1905,  p.  10,  March,  1905,  p.  16. 

383.  Cylinder  Lubrication.  —  The  test  requirements  for  cylinder  oils 
are  outlined  in  Table  89,  from  which  it  will  be  seen  that  pure  mineral 
oil  fulfills  practically  all  requirements  for  dry  steam.  In  connection 
with  moist  steam,  as  in  the  low-pressure  cylinders  of  compound  engines, 
an  addition  of  from  2  to  5  per  cent  of  acidless  tallow  oil  is  recom- 
mended. Vegetable  oils,  beeswax,  lard  oil,  degras  (wool  grease),  and 
the  like  should  never  be  used  in  compounding  cylinder  oils.  The  best 
cylinder  oils  are  made  from  Pennsylvania  stock. 

Cylinder  oils  must  be  forced  to  the  parts  requiring  lubrication 
against  the  prevailing  steam  pressure,  which  is  ordinarily  accomplished 
by  (1)  cylinder  cups,  (2)  hydrostatic  lubricators,  or  (3)  hand  or  power 
driven  force  pumps. 

383.  Cylinder  Cups.  —  A  cylinder  oil  cup  consists  essentially  of  a 
steam-tight  brass  vessel  fitted  at  the  bottom  with  a  pipe  connection 
and  valve.  A  screwed  cap  offers  a  means  of  introducing  the  lubri- 
cant into  the  cup.  After  the  cap  is  in  place  the  valve  is  opened  and 
the  cup  is  subjected  to  full  steam  pressure.  The  pressure  in  the  cup 
being  equal  to  that  in  the  steam  chest  or  cylinder,  permits  the  lubri- 
cant to  gravitate  through  the  valve  into  the  cylinder. 

Fig.  407  shows  a  section  through  an  improved  form  of  oil  cup  in 
which  the  oil  feeds  from  the  top  instead  of  the  bottom  as  is  the  case 
with  the  common  form  of  cylinder  cup.  The  vessel  is  attached  to  the 
steam  chest  or  to  the  supply  pipe  below  the  throttle  valve.  Steam  is 
admitted  through  opening  B  and,  condensing,  settles  through  the  oil 


LUBRICANTS  AND  LUBRICATION 


683 


to  the  bottom.  This  raises  the  level  of  the  oil  until  it  begins  to  over- 
flow down  the  same  passage  by  which  the  steam  enters.  This  action 
is  intensified  by  the  fluctuation  in  steam  pres- 
sure. The  rate  of  feeding  is  regulated  by  valve 
C  and  tested  by  unscrewing  plug  F.  If  oil 
appears  through  opening  G,  the  cup  is  feeding 
oil;  if  steam  or  water  is  emitted,  the  cup  is 
empty.  The  cup  is  filled  by  means  of  plug  E 
and  the  water  drained  at  Z>. 

384.  Hydrostatic  Lubricators.  —  The  most 
common  method  of  cylinder  lubrication  is  by 
means  of  hydrostatic  lubricators  of  the  sight-feed 
class,  Fig.  408.  The  principle  of  operation  is 
as  follows:  The  lubricator  is  filled  with  cylinder 
oil  by  removing  cap  K,  the  height  of  oil  ap- 
pearing in  glass  L.  If  water  is  present  the 
oil  floats  on  top  as  indicated.  After  the  cap 


FIG.  407.     Leyland  Auto- 
matic Cylinder  Cup. 


L/Q 

FIG.  408.    Common  Hydrostatic 
Lubricator. 


•central  reservoir. 


is  screwed  in  place  the  valves  in  the  con- 
denser pipe  are  opened,  subjecting  the 
oil  in  the  vessel  to  steam-pipe  pressure. 
Steam  is  condensed  in  pipe  C,  filling  tube 
B  and  part  of  C,  thus  adding  to  the  steam 
pressure  the  pressure  due  to  the  weight 
of  the  water  column.  Valve  F,  which 
communicates  with  the  top  of  the  vessel 
by  means  of  tube  A,  is  opened  wide,  as 
is  also  the  regulating  valve  /.  The  pres- 
sure at  B  being  greater  than  that  at  A 
by  an  amount  equivalent  to  the  height  of 
the  water  column,  forces  the  oil  through  A 
and  the  "  sight  feed  "  S  to  the  steam  pipe. 
The  rate  of  flow  is  controlled  by  the_regu- 
lating  valve  /.  As  the  oil  flows  from  the 
vessel  its  space  is  occupied  by  condensed 
steam,  the  height  of  oil  and  water  being 
visible  in  glass  L.  Owing  to  the  small 
capacity  of  the  lubricator  it  must  be  re- 
filled frequently.  To  reduce -%be-  amount 
of  labor  required  with  the  above  appa- 
ratus, independent  sight  feeds,  Fig.  409, 
are  sometimes  used  in  connection  with  a 
Such  an  installation  is  shown  diagrammatically  in 


684 


STEAM  POWER  PLANT  ENGINEERING 


Fig.  410.  A  condenser  pipe  leading  from  the  steam  main  enters  the 
bottom  of  the  reservoir  and  the  condensed  steam  fills  up  the  reservoir 

as  fast  as  the  oil  is  fed  out.  The  prin- 
ciple is  the  same  as  that  of  the  simple 
hydrostatic  lubricator.  Oil  is  frequently 
injected  by  mechanical  means  under  a 
steady  pressure  generated  and  governed 
independently  of  the  steam.  Two  sys- 
tems are  in  common  use,  direct  mechan- 
ical pump  pressure  and  air  pressure. 

385.  Forced-Feed  Cylinder  Lubrica- 
tion. —  Fig.  411  illustrates  the  "  Roches- 
ter" simple  feed  automatic  lubricating 
pump,  which  takes  the  oil  by  gravity 
from  the  reservoir  through  a  sight-feed 
glass  and  forces  it  through  a  small  pipe 
to  the  steam  supply  pipe.  The  pump 
entirely  obviates  the  trouble  due  to  in- 
termittent feeding  and,  being  directly 

FIG.  409.    Lunkenheimer  Sight-Feed     ,  .  -  ,, 

Lubricator.  driven  from   the    engine,    runs   at   con- 

stant  speed.     The   feed  is  uniform  and 

independent  of  the  pressure  pumped  against.  The  rate  is  deter- 
mined by  the  length  of  stroke  of  the  pump  piston,  which  is  easily 
adjusted. 


FIG.  410.    Central  Hydrostatic  Lubricator. 

With  large  engines  multi-feed  pumps    are    sometimes  used,  which 
force  oil  to  the  various  valves  as  well  as  to  the  steam  pipe.     Fig.  412 


LUBRICANTS  AND  LUBRICATION 


685 


shows  an  arrangement  of  storage  tank  in  connection  with  pump  reser- 
voir to  avoid  the  trouble  of  hand  filling. 


FIG.  411.     Rochester  Forced-Feed  Lubricator. 


H.P.Steam  Pipe 

UP.  Steam  Pipe          To  R^ 
To  Bod 


FIG.  412.    Forced-Feed  Cylinder  Lubrication. 

386.  Siegrist  System.  —  Fig.  413  shows  an  application  of  the 
Siegrist  system  of  cylinder  and  engine  lubrication.  There  are  two 
storage  tanks  on  the  engine-room  floor,  one  for  cylinder  oil  and  the 
other  for  engine  oil,  the  distributing  arrangements  being  the  same  in 
each  case.  The  oil  is  pumped  from  each  tank  into  a  main  pipe 
extending  the  length  of  the  engine  room  and  provided  with  branches 
at  each  point  requiring  lubrication.  The  oil  pumps  are  actuated  by 
steam  and  are  of  the  duplex  direct-acting  type,  provided  with  auto- 
matic governors  which  regulate  the  speed  to  suit  the  demand  for  oil. 


686 


STEAM  POWER  PLANT  ENGINEERING 


LUBRICANTS  AND  LUBRICATION 


687 


The  cylinder  oil  is  forced  through  a  special  sight-feed  lubricator,  Fig.  407, 
under  a  pressure  of  about  25  pounds  in  excess  of  the  steam  pressure. 
Referring  to  Fig.  414,  diaphragm  valve  D,  in  the  bottom  of  the  lubri- 
cator, is  kept  closed  by  the  steam  pressure  admitted  through  pipes  B. 
Thus  the  inlet  pressure  must  be  greater  than  that  of  the  steam  before 

the  valve  will  open  and  admit  oil 
to  the  engine.  The  oil,  after  enter- 
ing, passes  upward  through  the  sight- 
feed  glass  and  downward  through 
the  hollow  arm  A  to  the  steam  pipe. 
The  engine  oil  is  forced  by  the 
pump  to  the  various  points  under  a 
pressure  of  about  20  pounds.  The 
waste  oil  is  caught  in  suitable  re- 
ceptacles and,  after  being  filtered,  is 
returned  to  the  storage  tank  by  a 
steam  pump.  This  pump  is  con- 
nected so  that  it  can  supply  the 

FIG.  414.    Siegrist  Sight-Feed  Lubricator.     st°rag6  tank  elther  fr°m  the  fiiter  OI* 

with  fresh  oil  from  a  large  oil  tank 

in  the  basement.  By  this  arrangement  all  handling  of  oil  in  the  engine 
room  is  done  away  with. 

387.  Oil  Filters.  —  After  oil  has  been  applied  to  machinery  its 
lubricating  properties  become  impaired  on  account  of  (1)  contami- 
nation with  anti-lubricating  material,  such  as  dust,  metallic  particles 
from  wear,  gum,  acid,  and  resin;  and  (2)  exposure  to  heat  and  the 
atmosphere  which  drives  off  part  of  the  more  volatile  constituents  and 
decreases  the  fluidity  of  the  oil. 

In  many  small  plants  no  attempt  is  made  to  reclaim  oil  that  has 
once  been  used,  since  the  quantity  is  so  small  that  the  cost  and 
trouble  involved  would  more  than  offset  the  gain.  Where  large  quan- 
tities of  oil  are  used,  considerable  saving  may  be  effected  by  using 
it  over  and  over  again.  To  render  the  oil  fit  for  reuse  it  must  be 
thoroughly  purified.  The  anti-lubricating  matter  is  removed  by  pre- 
cipitation and  filtration. 

Fig.  415  shows  a  section  through  a  "  White  Star"  oil  filter  and  purifier. 
The  apparatus  consists  of  a  cylindrical  sheet-iron  vessel  divided  into  two 
compartments  by  a  vertical  partition.  These  two  compartments  are 
connected  near  the  top  by  valve  B.  The  smaller  chamber  is  provided 
with  a  funnel  A  and  a  steam  coil  for  heating  the  contents.  The  large 
chamber  contains  a  cylindrical  wire  screen  covered  with  several  folds  of 
filtering  cloth.  Impure  oil  is  poured  into  funnel  A,  the  upper  part  of 


688 


STEAM  POWER  PLANT  ENGINEERING 


WATER  LEVEL 


which  is  provided  with  a  removable  sieve  or  strainer,  and  is  discharged 
below  the  surface  of  the  water  through  holes  in  the  foot  of  the  tube. 
The  thin  streams  of  oil  rise  vertically  to  the  surface  of  the  water  and 
the  heavy  particles  of  grit  and  dirt  gravitate  to  the  bottom.  The  steam 

coil  heats  the  oil  and  water 
and  facilitates  precipita- 
tion of  the  solid  matter 
by  thinning  out  the  streams 
of  oil.  When  the  oil  in  the 
smaller  chamber  reaches  the 
level  of  valve  B  it  flows  in- 
to the  filter  bag,  which  re- 
moves the  remaining  im- 
purities and  permits  the 
purified  products  to  flow 
into  the  large  compartment 
from  which  it  may  be  drawn 
at  will.  All  parts  are  access- 
ible and  readily  removed 
for  cleaning  purposes.  The 
accumulated  sediment  in  the  bottom  of  the  small  chamber  is  dis- 
charged to  waste  at  intervals  by  means  of  a  suitable  drain.  When  the 


STRAINER 

PERFORATED  PLATE 
FILTERING  MATERIAL 
PERFORATED  PLATE 

PERFORATED  PLATE 
FILTERING  MATERIAL 

PERFORATED  PLATE 
WATER 

STEAM  COILS 


FIG.  415.    White  Star  Oil  Filter. 


SECTION  1  SECTION     2  SECTION     3 

FIG.  416.    Turner  Oil  Filter. 


SECTION   4 


filter  cloth  is  to  be  removed,  valve  B  is  closed  and  the  wire  cylinder  is 
disconnected  and  lifted  out.  Any  oil  remaining  in  the  filter  is  returned 
to  funnel  A.  The  filter  cloth  is  held  against  the  screen  by  cords  and 
hence  is  readily  removed. 


LUBRICANTS  AND  LUBRICATION  689 

Fig.  416  shows  a  section  through  a  Turner  oil  filter,  illustrating  the 
type  of  filter  usually  installed  in  large  stations  where  continuous  fil- 
tration is  desired.  This  apparatus  consists  of  a  rectangular  tank 
divided  into  four  compartments.  The  returns  from  the  lubricating 
system  flow  into  section  1  through  a  screened  funnel  and  discharge 
into  the  water  space  at  the  bottom  of  the  compartment.  The  oil  rises 
through  the  water,  passes,  under  pressure  of  the  head  in  the  funnel, 
through  a  layer  of  filtering  material  resting  on  a  perforated  plate,  and 
collects  in  an  inverted  cone.  Through  perforations  round  the  top  of 
the  cone  it  passes  into  a  dirt  chamber,  where  most  of  the  heavy  impuri- 
ties are  deposited,  and  then,  still  rising,  passes  through  another  per- 
forated plate  and  more  filtering  material.  The  partially  cleaned  oil, 
which  issues,  overflows  into  the  second  compartment  and  thence  into 
the  third,  the  same  cycle  of  operations  being  repeated  in  these  two. 
The  overflow  from  the  third  compartment  descends  through  a  final 
filter  in  the  fourth  compartment  and  collects  at  the  bottom,  from 
which  it  is  withdrawn  by  the  oil  pump. 

Forced-Feed  Lubrication:  Am.  Elecn.,  Aug.,  1902,  p.  402,  Dec.,  1905,  p.  608; 
Automobile,  Nov.  1,  1906,  p.  572;  Mech.  Engr.,  April  20,  1907,  p.  552. 

Cylinder  Lubrication:  Power,  Dec.,  1902,  p.  30,  Jan.,  1905,  p.  36,  March,  1906, 
p.  163;  Engr.,  Lond.,  1905,  Vol.  96,  pp.  55,  108,  132,  155;  St.  Ry.  Jour.,  June  22,  1907, 
p.  1103;  Engr.  U.S.,  Oct.  15,  1906,  p.  682;  Am.  Gas  Light  Jour.,  Jan.  23,  1905,  p.  130; 
Horseless  Age,  Sept.  24,  1902,  p.  676. 

Miscellaneous. — Measurement  of  Durability  of  Lubricants:  Trans.  A.S.M.E., 
11-1013.  Valuation  of  Lubricant  by  Consumer:  Trans.  A.S.M.E.,  6-437.  Suit- 
ability of  Lubricants:  Power,  Nov.,  1906,  p.  673.  Oil  Required  for  Lubricators: 
Elec.  World,  May  5,  1906,  p.  934.  Gumming  Tests:  Jour.  Am.  Chem.  Soc.,  April, 
1902,  p.  467.  Valuation  of  Lubricants:  Jour.  Am.  Chem.  Ind.,  April  15,  1905, 
p.  315. 

Lubrication,  General:  Power,  March,  1903,  p.  135;  Mech.  Engr.,  June  30,  1906, 
p.  919;  Prac.  Engr.,  Dec.  15,  1905,  p.  915;  Elec.  Engr.,  Lond.,  Sept.  7,  1906,  p.  344. 

Oil  Purification:  Elec.  Engr.,  Lond.,  Jan.  13,  1903,  p.  51;  Elec.  World,  Dec. 
1,  1906,  p.  1053. 

Economy  in  Lubrication  of  Machinery:  Trans.  A.S.M.E.,  4-315.  Theory  of 
Finance  of  Lubrication:  Trans.  A.S.M.E.,  6-437. 

Experiments,  Formulas,  and  Constants  for  Lubrication  of  Bearings:  Am.  Mach., 
Sept.  10,  1903,  pp.  1281,  1316,  1350. 

Lubricators  and  Lubricants:  Power  &  Engr.,  Sept.  21,  1909,  p.  486.  / 

Selection  of  an  Oil  for  Lubrication:  Power  &  Engr.,  July  27,  1909,  p.  137. 


CHAPTER  XVII. 

FINANCE  AND   ECONOMICS  — COST   OF   POWER. 

388.  Records.  —  Few  engineers  realize  the  importance  of  a  detailed 
system  of  accounting,  or  the  saving  which  may  be  effected  in  cost  of 
operation  by  careful  study  of  the  daily  records  of  performance.     Many 
regard  graphical   load   curves,  meter  readings,  and  similar  records  as 
interesting  but  of  little  economic  value.      During  the  past  few  years 
the  author  has  made  a  close  study  of  the  cost  of  power  in  a  large  num- 
ber of  central  and  isolated  stations  in  Chicago,  and  found,  without 
exception,  that  the  highest  economy  was  effected  by  the  engineers 
who    kept    the    most    systematic   records;    the   poorest   results    were 
obtained  where  records  were  kept  indifferently  or  not  at  all.     In  some 
small  plants  the  numerous  duties  of  the  engineer  prevented  him  from 
devoting  the  necessary  time,  but  in  the  majority  of  cases  the  absence 
of  records  was  due  entirely  to  lack  of  interest.     Power-plant  records  to 
be  of  value  must  be  closely  studied  with  a  view  to  improvement.     The 
mere  accumulation  of  data  to  be  filed  away  and  never  again  referred 
to  is  a  waste  of  time  and  money. 

Records  should  cover  not  only  the  daily  operation  of  the  plant  but 
also,  as  permanent  statistics,  a  complete  analysis  of  each  item  of 
equipment.  The  value  of  such  data  cannot  be  overestimated.  The 
engineer  will  frequently  find  it  greatly  to  his  interest  to  have  avail- 
able at  a  moment's  notice  the  complete  details  of  his  engines,  boilers, 
generators,  and  other  machinery,  especially  when  it  is  required  to 
renew  a  broken  or  worn-out  part. 

389.  Output.  —  The  periodical  output  of  a  power  plant   may  be 
expressed  in  terms  of 

(1)  Steam  plant. 

Indicated  or  brake  horse  power. 
Indicated  or  brake  horse-power  hours. 

(2)  Electric  lighting  plant. 

Electrical  horse  power  or  kilowatts. 
Electrical  horse-power  hours  or  kilowatt  hours. 
Lamp  hours. 

690 


FINANCE  AND  ECONOMICS  —  COST  OF  POWER.  691 

(3)  Electric  railway  plant. 
Electrical  horse  power  or  kilowatts. 
Electrical  horse-power  hours  or  kilowatt  hours. 
Car  miles. 

When  a  plant  is  operating  at  practically  constant  load  it  is  suffi- 
ciently accurate  for  most  purposes  to  express  the  output  in  horse 
power  or  kilowatts  per  year.  When  the  output  fluctuates  from  day 
to  day  it  is  best  expressed  in  horse-power  hours  or  kilowatt  hours,  or 
by  specifying  the  load  factor  along  with  the  periodical  output  in 
horse  power.  For  example,  1  horse  power  per  year,  24  hours  per  day 
and  365  days  per  year,  is  equivalent  to  365  X  24  =  8760  horse-power 
hours.  If  the  full  power  is  used  throughout  this  time,  it  matters  little 
whether  the  charge  is  based  on  horse  power  or  horse-power  hours;  if, 
however,  the  power  is  used  say  only  half  the  time,  the  yearly  cost  per 
horse  power  will  remain  unchanged  but  the  cost  per  horse-power  hour 
will  be  just  double.  As  will  be  shown  later  the  load  factor  (ratio  of 
actual  to  rated  load)  exerts  a  marked  influence  on  the  cost  of  pro- 
ducing power,  and  for  this  reason  the  output  is  usually  expressed  as 
horse-power  hours,  kilowatt  hours,  lamp  hours,  or  the  like. 

390.  Load  Factor.  —  The  yearly  load  factor  or  simply  load  factor,  as 
it  is  usually  called,  is  the  ratio  of  the  actual  yearly  output  to  the  rated 
yearly  output  measured  on  a  24-hour  basis.  Thus: 

For  a  steam  plant, 

Load  factor  =  Yearly  output,  horse-power  hours ;  ( 

Rated  horse  power  X  8760 

For  an  electric  station, 

Load  factor  =     Yearly  output,  kilowatt  hours  ( 

Rated  capacity,  kilowatts,  X  8760 
(8760  =  number  of  hours  in  one  year.) 

The  curve  load  factor  or  station  load  factor  is  the  ratio  of  the  actual 
yearly  output  to  the  rated  output,  based  upon  the  number  of  hours  the 
plant  is  in  actual  operation.  Thus  for  an  electric  station, 

Curve  load  factor  = Yearly  output,  kilowatt  hours 

Rated  capacity  X  hours  plant  is  in  operation 

In  any  plant  the  great  desideratum  is  a  high  load  factor  with  con- 
sequent greatest  return  on  the  investment.  All  the  factors  of  expense 
included  in  the  cost  of  power  are  then  operating  at  maximum  economy. 
High  peak  loads  and  low  average  loads  necessitate  large  machines 
which  are  but  little  used  and  greatly  increase  the  fixed  charges. 


692 


STEAM  POWER  PLANT  ENGINEERING 


In  any  system  the  total  fixed  charges  per  year  are  constant  irre- 
spective of  the  load  factor,  since  interest,  taxes,  depreciation,  insurance, 
and  maintenance  go  on  whether  the  plant  is  in  operation  or  not.  The 
total  fixed  charges  for  a  specific  case  are  illustrated  in  Fig.  417  by  a 


320000 


30  40  50  60  70 

Yearly  Load  Factor— Per  Cent 


80 


100 


FIG.  417. 


Influence  of  Load  Factor  on  the  Cost  of  Power  at  the  Switchboard. 
Kilowatt  Electric  Light  and  Power  Station.) 


(5000 


straight  line.  The  cost  per  kilowatt  hour,  however,  decreases  as  the 
load  factor  increases.  For  example,  with  the  plant  operating  con- 
tinuously at  rated  load  (100  per  cent  load  factor)  the  fixed  charges 
per  kilowatt  hour  are 

65,000 


5000  X  8760 


$0.00148. 


With  30  per  cent  load  factor  these  charges  are 


65,000 


0.3  (5000  X  8760) 


$0.00445  kilowatt  hour. 


The  higher  the  load-factor  the  greater  is  the  amount  of  power  produced 
and  the  longer  does  the  apparatus  work  at  best  efficiency.  But  the 
greater  the  power  produced  the  larger  will  be  the  fuel  consumption 
and  the  oil  and  supply  requirements.  The  labor  charges  will  be  prac- 
tically constant.  The  total  operating  cost  per  year  increases  as  the 
load  factor  increases,  but  not  directly.  (See  Fig.  417.)  The  cost  per 


FINANCE  AND  ECONOMICS  —  COST  OF  POWER  693 

kilowatt  hour,  however,  decreases  as  the  load  factor  increases.  For 
example,  the  operating  costs  per  year  with  plant  operating  contin- 
uously at  full  load  are  $230,200.  This  gives 

23Q?20Q —  =  $0.00525  per  kilowatt  hour. 
5000  X  8760 

With  30  per  cent  load  factor  the  yearly  operating  charges  are 
$87,980,  which  gives 

87,980 =  $o.0067  per  kilowatt  hour. 

0.3  (5000  X  8760) 

Table  107  shows  the  influence  of  the  load  factor  on  the  cost  of 
power  in  two  isolated  stations  of  the  same  rated  capacity,  one  operat- 
ing with  the  unusually  high  load  factor  of  80  per  cent  and  the  other 
operating  with  the  low  load  factor  of  17  per  cent.  The  former  fur- 
nishes current  for  a  large  electro-chemical  concern  in  which  the  load 
is  practically  constant. 

In  general,  the  higher  the  load  factor  the  greater  becomes  the  ratio 
of  the  operating  to  the  fixed  charges,  and  extra  investment  may  become 
advisable  to  secure  the  greatest  economy  possible. 

On  the  other  hand,  when  the  load  factor  is  low  the  fixed  charges  are 
the  governing  factor  in  the  cost  of  power,  and  extra  expenditures 
must  be  carefully  considered,  particularly  if  fuel  is  cheap. 

391.  Cost  of  Operation.  —  The  cost  of  operation  of  power  plants  is 
conveniently  divided  into  two  parts: 

(1)  Fixed  charges. 

(a)  Investment  costs. 

(b)  Administration  costs. 

(2)  Operating  costs. 

392.  Fixed    Charges.  —  These    cover    all    expenses    which    do    not 
expand  and  contract  with  the  output.     In  very  large  plants  they  are 
usually  divided  into  two  parts,  (a)  the  investment  costs,  which  include 
interest,  rental,  depreciation,  taxes,  and  insurance,  and  a  reserve  fund 
to  cover  depreciation  of  the  investment,  and  (b)  the  administration 
costs,  which  include  rental  of  offices,  annual  salaries  of  officers,  and  all 
other  expenses  not  directly  chargeable  to  the  power  plant.     In  the 
average  plant  the  fixed  charges  comprise  interest,  rental,  depreciation, 
taxes,  insurance,  and  sometimes   maintenance,   though  the  latter  is 
ordinarily  included  in  the-  operating  costs. 

393.  Interest.  —  The   rates   of   interest   on   borrowed   money   vary 
with  the  nature  of  the  security.     In  the  case  of  power  plants  the  form 
of  security  is  usually  a  mortgage  on  the  plant  and  equipment.     If  a 


694  STEAM  POWER  PLANT  ENGINEERING 

builder  has  sufficient  funds  to  construct  the  plant  without  borrowing, 
he  should  charge  against  the  item  "  interest  "  the  income  which  the  sum 
involved  would  bring  if  placed  out  at  interest  or  if  invested  in  his  own 
business.  In  estimating  the  interest  charges  5  per  cent  of  the  capital 
invested  is  ordinarily  assumed  unless  specific  figures  are  available. 

TABLE  90. 

APPROXIMATE  USEFUL  LIFE  OF  VARIOUS  PORTIONS  OF  STEAM  POWER  PLANT 

EQUIPMENTS. 

Years. 

Buildings,  brick  or  concrete 50 

Buildings,  wooden  or  sheet  iron 15 

Chimneys,  brick 50 

Chimneys,  self-sustaining  steel 25 

Chimneys,  guyed  sheet-iron 10 

Boilers,'  water-tube 25 

Boilers,  fire-tube 15 

Engines,  slow-speed 25 

Engines,  high-speed 15 

Turbines : 25 

Generators,  direct-current 25 

Generators,  alternating-current 30 

Motors 20 

Pumps 25 

Condensers,  jet 35 

Condensers,  surface 20 

Heaters,  open 30 

Heaters,  closed 20 

Economizers 20 

Wiring 20 

Belts 7 

Coal  conveyor,  bucket 15 

Coal  conveyor,  belt 10 

Transformers,  stationary 30 

Rotary  converters 25 

Storage  batteries 15 

Piping,  ordinary 12 

Piping,  first  class 20 

NOTE. —  So  much  depends  upon  the  design  and  the  conditions  of  operation  that  no  fixed 
values  can  be  definitely  assigned  and  the  above  figures  should  be  used  with  caution.  Practice 
shows  that  most  power-plant  appliances  become  obsolete  long  before  the  limit  of  their  useful 
life  is  reached. 

394.  Depreciation.  —  This  charge  represents  the  gradual  deterio- 
ration of  a  plant,  resulting  in  its  eventually  wearing  out.  It  is  also 
assumed  to  represent  the  superannuation  of  a  plant  or  the  rate  at 
which  the  apparatus  is  becoming  obsolete.  Thus,  under  the  first 
assumption,  if  the  useful  life  of  an  engine  is  40  years,  the  rate  of 
depreciation,  neglecting  interest,  is  2.5  per  cent;  if,  however,  it  is 
assumed  that  the  engine  will  become  obsolete  in  20  years  and  uneco- 
nomical for  further  operation,  the  rate  of  depreciation  will  be  5  per 
cent.  It  is  difficult  to  assign  a  fixed  rate  of  depreciation  against  any 


FINANCE  AND  ECONOMICS  —  COST  OF  POWER 


695 


piece  of  apparatus,  due  to  possible  new  developments  which  cannot 
be  reckoned  with  in  advance  in  computing  the  useful  life  of  the  appa- 
ratus. Again,  depreciation  cannot  always  be  separated  from  current 
repairs  and  is  a  variable  factor  even  in  the  parts  of  the  same  machine. 
It  is  therefore  more  or  less  of  an  approximation.  The  average  life  of 
various  parts  of  a  steam  power  plant  is  outlined  in  Table  90,  but 
on  account  of  the  inability  to  assign  fixed  values  to  the  useful  life  of 
any  apparatus,  and  on  account  of  the  great  number  of  appliances  in 
even  a  small  plant,  it  is  customary  to  charge  a  fixed  rate  of  depreciation 
against  the  entire  plant  and  thus  avoid  confusion  and  complexity. 
This  very  crude  method  usually  results  in  overestimation  in  well- 
designed,  well-operated  plants  and  underestimation  in  poorly  designed 
and  badly  managed  installations.  One  of  the  largest  power  plant 
designing  concerns  in  Chicago  charges  7J  per  cent  against  deprecia- 
tion and  finds  this  figure  none  too  small.  The  Pennsylvania  Railroad 
uses  7  per  cent  to  cover  depreciation  charges  on  their  power-house 
equipments. 

TABLE  91. 

RATE    OF    DEPRECIATION. 
(Per  Cent  of  First  Cost.) 


Rate  of  Interest,  per  Cent. 

2 

2.5 

3 

8.5 

4 

4.5 

5 

5.5 

6 

7 

8 

9 

10 

2 

49.50 

49.37 

49.27 

49.14 

49.02 

48.90 

48.78 

48.66 

48.54 

48.31 

48.07 

47.84 

47.62 

3 

32.67 

32.51 

32.35 

32.19 

32.03 

31.87 

31.72 

31.56 

31.41 

31.10 

30.80 

30.51 

30.21 

4 

24.26 

24.08 

23.90 

23.72 

23.55 

23.39 

23.20 

23.03 

22.86 

22.52 

22.19 

21.84 

21.55 

5 

19.21 

19.02 

18.83 

18.65 

18.46 

18.28 

18.10 

17.91 

17.73 

17.40 

17.04 

16.73 

16.37 

6 

15.85 

15.65 

15.46 

15.26 

15.08 

14.89 

14.70 

14.52 

14.33 

13.97 

13.63 

13.29 

12.96 

3 

7 

13.45 

13.25 

13.05 

12.85 

12.66 

12.46 

12.28 

12.09 

11.91 

11.15 

11.20 

10.87 

10.55 

Z3 
ffl 

8 

11.65 

11.44 

11.24 

11.05 

10.85 

10.66 

10.47 

10.28 

10.10 

9.74 

9.40 

9.06 

8.74 

9 

10.25 

10.04 

9.84 

9.64 

9.45 

9.26 

9.07 

8.88 

8.70 

8.34 

8.00 

7.68 

7.36 

3! 

10 

9.13 

8.92 

8.72 

8.52 

8.33 

8.14 

7.95 

7.76 

7.58 

7.23 

6.90 

6.58 

6.27 

•< 

11 

8.21 

8.01 

7.80 

7.61 

7.41 

7.22 

7.04 

6.85 

6.68 

6.33 

6.00 

5.69 

5.40 

o 

12 

7.45 

7.25 

7.04 

6.85 

6.65 

6.46 

6.28 

6.10 

5.92 

5.60 

5.27 

4.97 

4.69 

9 

13 

6.81 

6.60 

6.40 

6.20 

6.01 

5.83 

5.64 

5.47 

5.29 

4.96 

4.65 

4.36 

4.08 

3 

14 

6.26 

6.05 

5.85 

5.65 

5.46 

5.28 

5.10 

4.93 

4.75 

4.49 

4.13 

3.84 

3.58 

1 

15 

5.78 

5.57 

5.37 

5.18 

4.99 

4.81 

4.63 

4.46 

4.29 

3.97 

3.68 

3.40 

3.15 

^  - 

16 

5.36 

5.16 

4.96 

4.77 

4.58 

4.40 

4.22 

4.06 

3.89 

3.58 

3.30 

3.03 

2.78 

17 

4.99 

4.79 

4.59 

4.40 

4.22 

4.04 

3.87 

3.70 

3.54 

3.24 

2.96 

2.71 

2.47 

$ 

| 

18 

4.67 

4.46 

4.27 

4.08 

3.90 

3.72 

3.55 

3.39 

3.23 

2.94 

2.66 

2.42 

2.19 

19 

4.37 

4.17 

3.98 

3.79 

3.61 

3.44 

3.27 

3.11 

2.96 

2.67 

2.47 

2.17 

1.95 

** 

20 

4.11 

3.91 

3.72 

3.53 

3.36 

3.19 

3.02 

2.87 

2.71 

2.44 

2.18 

1.95 

1.95 

25 

3.12 

2.92 

2.74 

2.56 

2.40 

2.24 

2.09 

1.95 

1.82 

1.58 

1.36 

1.18 

1.75 

30 

2.46 

2.27 

2.10 

1.93 

1.78 

1.64 

1.50 

1.38 

1.26 

1.06 

0.88 

0.73 

0.61 

35 

2.00 

1.82 

1.65 

1.50 

1.36 

1.23 

1.10 

0.99 

0.89 

0.72 

0.58 

0.46 

0.37 

40 

1.65 

1.48 

1.32 

1.18 

1.05 

0.93 

0.83 

0.73 

0.64 

0.50 

0.38 

0.29 

0.22 

45 

1.39 

1.22 

1.07 

0.94 

0.82 

0.72 

0.62 

0.54 

0.47 

0.35 

0.26 

0.19 

0.14 

50 

1.18 

1.02 

0.88 

0.76 

0.65 

0.56 

0.42 

0.40 

0.34 

0.25 

0.17 

0.12 

0.09 

696  STEAM  POWER  PLANT  ENGINEERING 

The  rate  of  depreciation  in  terms  of  interest  and  useful  life  is  a 
simple  problem  in  compound  interest,  and  may  be  expressed: 


in  which  r     ' 

d  =rate  of  depreciation,  per  cent  of  first  cost. 

r  =  rate  of  interest. 

n  =  assumed  life  of  the  apparatus  in  years. 

This  is  based  on  the  assumption  that  the  interest  is  compounded 
annually  and  that  the  apparatus  is  valueless  at  the  end  of  n  years. 
Table  91  has  been  calculated  with  this  formula  and  gives  the  rate  of 
depreciation  for  different  rates  of  interest  and  different  asssumptions 
as  to  useful  life  of  apparatus. 

TABLE   92. 

DEPRECIATION    PERCENTAGES    DETERMINED    BY    THE    TRACTION    VALUATION 
COMMISSION.    POWER-PLANT    DEPRECIATION. 

Chicago,  111.,  Sept.  8,  1906.  Per  Cent. 

Engines,  Corliss,  slow-speed  ......................................  3  to  5 

Engines,  automatic,  high-speed  ..................................  5  to  10 

Cable-winding  machinery  ...........................................   3 

Generators,  direct  connected,  modern  ................................  5 

Generators,  belted  (depending  on  date)  ..........................  5  to  10 

Traveling  cranes  .....................................................  2 

Switchboard  and  all  wiring  ...........................................  2 

Piping  ...............................................................  35 

Pumps.  .  .  ".  ...........................................................  5 

Heaters,  closed  .................................................  6  to  10 

Heaters,  open,  if  cast  iron  only  .......................................  3 

Breeching  and  connections,  brick  ....................................  5 

Breeching  and  connections,  steel  ....................................  10 

Boilers  and  settings,  horizontal  tubular  ..............................  10 

Boilers  and  settings,  water-  tube  ....................................  3.5 

Grates  ..............................................................  10 

Stokers  ..........................  -.  .........................    See  below 

Coal-handling  machinery  .............................................  6 

Ash-handling  machinery  ..............................................  & 

Combined  coal  and  ash-handling  machinery  ..........................  7 

Storage  bins,  steel  ..............................................  3  to  10 

Miscellaneous  items  ..................................................  5 

The  above  annual  rates  of  depreciation  have  been  used  as  a  basis  in  depreciating 
the  power-plant  equipments.  Apparatus  has  been  depreciated  at  these  rates  down 
to  20  per  cent  of  the  wearing  value,  the  wearing  value  being  determined  by  sub- 
tracting the  scrap  value  from  the  cost  new.  All  power-plant  equipment  has  been 
considered  as  worth  20  per  cent  of  its  wearing  value  as  long  as  it  is  in  operating 
condition.  Depreciation  applied  to  wearing  value,  as  apparatus  is  always  worth 
scrap  value. 

Stokers.  The  fixed  parts  depreciate  very  little  and  the  moving  parts  and  grates 
very  rapidly,  as  the  moving  parts  are  renewed  and  maintained  in  good  condition. 
All  stokers  in  operation  have  been  depreciated  25  per  cent. 

The  above  percentages  applied  to  a  particular  plant  of  3900  kilowatts  capacity 
give  an  approximate  depreciation  for  the  whole  plant  of  4  per  cent. 


FINANCE  AND  ECONOMICS  —  COST  OF  POWER  697 

Table  92  gives  the  depreciation  percentages  determined  by  the 
Traction  Valuation  Commission,  Chicago,  111.,  as  reported  by  the  com- 
mission Sept.  18,  1906. 

Example:  A  1000-square-feet  surface  condenser  and  auxiliaries  cost 
$3500.  With  interest  at  5  per  cent,  required  the  rate  of  depreciation, 
assuming  a  life  of  25  years. 


(l+  0.05     -l 

That  is  to  say,  if  2.09  per  cent  of  the  first  cost  is  laid  aside  each  year 
for  25  years  at  5  per  cent  interest,  compounded  annually,  it  will  equal 
the  first  cost  at  the  end  of  this  period.  The  sum  thus  laid  aside 
is  sometimes  called  the  sinking  fund.  The  solution  is  more  readily 
obtained  with  the  aid  of  Table  91;  e.g.,  at  the  intersection  of  vertical 
column  5  (interest)  and  horizontal  column  25  (life  in  years)  we  find 
the  depreciation  2.09  per  cent.  This  sinking-fund  method  of  rating 
the  depreciation  is  peculiarly  adapted  to  power-plant  practice,  inas- 
much as  a  sum  set  aside  at  comparatively  low  rates  of  interest  and 
compounded  increases  very  slowly  at  first  but  more  and  more  rapidly 
from  year  to  year.  This  is  precisely  what  happens  in  the  deterioration 
of  the  plant.  The  loss  of  value  is  slight  at  first  when  the  materials 
are  new  and  usefulness  is  at  a  maximum,  while  towards  the  end  of 
life  both  value  and  usefulness  decline  very  rapidly. 

If  the  apparatus  still  has  some  value  at  the  end  of  n  years  and  if  b  is 
the  ratio  of  the  value  of  the  old  material  to  that  of  the  new,  the  rate 
of  depreciation  becomes 

d  =  100     r(1~6)  —  (174) 

(1  +  r)»-l  v 

Example  :  In  the  foregoing  problem,  required  the  rate  of  depreciation 
if  the  value  of  the  condenser  outfit  is  $350  at  the  end  of  25  years. 


d  =    100  X  0.05  (1-0.1) 
(1  +  .05)25  -  1 

==  1.97  per  cent. 

That  is,  1.97  per  cent  of  $3500,  or  $68.95,  laid  aside  each  year  for 
25  years  at  5  per  cent  interest  and  compounded  annually  will  equal 
$3500  —  350,  or  $3150,  at  the  end  of  this  period. 


698  STEAM  POWER  PLANT  ENGINEERING 

It  is  not  supposed  that  an  owner  will  regularly  lay  aside  this  sum 
annually,  or  take  the  trouble  to  arrange  for  its  investment  at  current 
rates  in  the  market  or  savings  bank,  since  the  money  is  probably  worth 
more  to  him  in  his  own  business.  In  practice  it  is  retained  in  his 
business  or  investments  and  is  earning  the  rate  of  interest  obtainable 
therein,  but  in  determining  the  net  profit  or  loss  this  depreciation  item 
is  nevertheless  accounted  for  just  as  if  it  were  actually  placed  in  out- 
side investments. 

In  appraising  the  present  value  of  any  apparatus  in  terms  of  the  rate 
of  interest  and  useful  life  the  expression  becomes 


__        

100     r--  <175> 


in  which 


V  =  total  depreciation,  per  cent  of  original  cost. 

m  =  number  of  years  apparatus  has  been  in  operation. 

n  =  assumed  life  of  apparatus. 

r  =  rate  of  interest. 

Example:   In  the  preceding  example,  required  the  present  valuation 
of  the  condenser,  assuming  that  it  has  been  in  use  15  years. 

m  =  15,  n  =  25,  r  =  0.05. 
Substituting  these  values  in  (175), 


That  is,  the  "condenser  has  depreciated  45.1  per  cent  of  its  original 
value  and  consequently  is  worth  $3500  —  0.451  X  3500  =  $1921.50. 

Table  91  may  be  conveniently  used  in  'this  connection.  At  the 
intersection  of  vertical  column  5  and  horizontal  columns  15  and  25 
we  find  4.63  and  2.09  respectively.  Dividing  2.09  by  4.63  we  get 
0.451  =  45.1  per  cent,  the  total  depreciation.  Depreciation  is  often 
taken  care  of  under  the  different  items  pertaining  to  maintenance,  and 
whenever  a  change  or  repair  is  necessary  it  is  charged  directly  into 
expense  as  maintenance. 

Though  usually  considered  separately,  interest  and  depreciation  are 
sometimes  considered  as  a  single  item,  and  in  this  case  the  rate  of 
depreciation  represents  the  sum  which  must  be  laid  aside  each 
year  for  the  eventual  renewal  of  apparatus  plus  the  interest  on  the 
investment. 


FINANCE  AND  ECONOMICS  -  COST  OF  POWER  699 

395.  Maintenance.  —  Maintenance  usually  refers  to  the  expense  of 
keeping  the  plant  in  running  order  over  and  above  the  cost  of  attend- 
ance.    It  includes  cost   of    upkeep,  replacement,   and  precautionary 
measures.     This   latter  item   includes   the  renewal  of  working  parts, 
painting  of  perishable  or  exposed  material,  and  replacing  worn-out  and 
defective  material.     Many  engineers  make  no  allowance  for  mainte- 
nance in  the  fixed  charges   and  include  these  costs  under  supplies, 
attendance,    or    repairs.     In    a    general    way,    when    maintenance    is 
included  under  the  fixed  charges,  an  annual  charge  of  2  per  cent  is  con- 
sidered a  liberal  allowance,  since  most  of  the  repair  work  comes  under 
attendance.     In  street-railway  practice  maintenance  is  divided  among 
the  several  parts   of   the  system  as  follows:   Buildings,  steam  appli- 
ances, electrical  equipment,  and  miscellaneous.    In  this  connection  the 
maintenance  becomes  a  part  of  the  fixed  charges,  since  the  various 
items  vary  widely  from  month  to  month. 

396.  Taxes  and  Insurance.  —  Taxes  vary  from  a  fraction  of  one  per 
cent  to  one  and  one-half  per  cent,  depending  upon  the  location  of  the 
plant.     An  average  figure  is  one  per  cent  of  the  actual  value  of  the 
investment.     Buildings  and  machinery  are  ordinarily  insured  against 
fire  loss  and  boilers  against  accidental  explosions,  and  accident  policies 
are  sometimes  carried  on  all  operating  machinery.     A  fair  charge  for 
this  item  is  one-half  per  cent. 

397.  Operating  Costs.  —  Operating  costs  are  conveniently  divided  as 
follows : 

(1)  Labor  or  attendance. 

(2)  Fuel  and  water. 

(3)  Oil,  waste,  and  supplies. 

(4)  Repairs  and  maintenance. 

In  large  stations  it  is  often  desirable  to  keep  the  expenses  of  the 
various  departments  separate  from  those  of  the  power  plant  proper. 
Thus  in  central  stations  the  distributing  system  is  an  important 
branch  and  the  attending  expenses  form  a  considerable  portion  of  the 
total.  They  are  therefore  kept  separate,  si$ce  they  are  not  strictly 
chargeable  to  power  generation.  In  isolated  stations  the  wages  of 
elevator  men,  though  in  a  general  way  a  part  of  the  power-plant 
expenses,  are  not  included  in  the  "  labor  and  attendance  "  charge  of 
the  plant.  Lamps  are  a  large  item  of  expense  in  a  tall  office  build- 
ing, and  for  this  reason  are  often  kept  separate  from  supplies. 

398.  Labor,  Attendance,  Wages.  —  The  minimum  number  of  men 
required  to  handle  a  given  plant  is  approximately  a  fixed  quantity 
and  it  is  seldom  possible  to  so  arrange  the  work  that  any  material 


700  STEAM  POWER  PLANT  ENGINEERING 

reduction  can  be  effected.  Until  very  recently  it  has  been  the  uni- 
versal custom  to  pay  wages  on  a  "  flat  rate  "  basis,  that  is,  the  attend- 
ant is  given  a  fixed  sum  per  day  or  month  irrespective  of  the  amount 
of  work  required  or  the  economy  of  operation.  In  many  cases,  how- 
ever, the  bonus  system  has  been  successfully  adopted.  For  example, 
in  the  boiler  room  the  coal  consumption  is  determined  for  a  given 
period  of  time  with  ordinary  careful  firing,  and  the  fireman  is  offered 
a  reasonable  percentage  on  the  saving  of  coal  which  he  is  able  to 
effect  over  this  record  by  special  care  and  attention  to  the  keeping  of 
fires  always  in  the  best  condition,  avoiding  the  blowing  off  of  steam, 
using  as  little  coal  as  needed  for  banking  fires,  and  in  other  ways. 
Where  careful  records  are  kept  of  supplies,  repairs,  and  renewals,  the 
bonus  is  also  applicable  to  electricians,  oilers,  and  other  employees. 

Labor  should  always  be  estimated  or  recorded  as  so  many  dollars 
per  month  or  per  year  and  not  merely  in  terms  of  the  output  unless 
the  load  factor  is  definitely  known,  otherwise  comparisons  are  mis- 
leading. For  example,  consider  two  plants  of  500  kilowatts  capacity, 
each  with  labor  charges,  say,  of  $400  per  month.  Suppose  the  output 
of  one  is  100,000  kilowatt  hours  per  month  and  that  of  the  other 
40,000  kilowatt  hours  per  month.  The  monthly  charges  are  evidently 
the  same,  viz.,  $400,  but  the  cost  per  kilowatt  hour  differs  widely, 
being  0.4  cent  in  the  first  case  and  1  cent  in  the  latter. 

The  cost  of  labor  varies  so  much  with  the  location  of  the  plant  and 
the  conditions  of  operation  that  general  figures  are  of  little  value 
except  as  a  rough  guide.  The  figures  in  Table  93  and  Table  94  give 
average  results  for  general  practice.  Specific  figures  will  be  found  in 
Tables  95  to  107. 

399.  Cost  of  Fuel.  —  Tables  95  to  107  give  examples  of  the  cost  of 
fuel  in  different  types  and  sizes  of  steam  power  plants.  It  will  be 
noted  that  this  item  varies  considerably  even  with  plants  of  the  same 
general  class.  So  much  depends  upon  the  market  price  of  the  fuel 
itself  that  the  item  "  cost  per  horse-power  or  kilowatt  hour  "  gives 
little  information  concerning  the  economy  of  operation  unless  the 
price  of  the  fuel,  its  calorific  value,  and  the  water  rate  of  the  prime 
movers  are  specified.  In  a  general  sense  the  cost  of  fuel  will  range 
from  40  to  70  per  cent  of  the  total  operating  expenses.  In  estimating 
the  cost  of  fuel  for  a  proposed  installation  due  consideration  should  be 
given  to  the  coal  burned  in  banking  fires,  heat  lost  in  blowing  off 
boilers,  and  reduced  efficiency  in  operating  at  underloads  and  over- 
loads. For  example,  individual  tests  of  a  number  of  boilers  in  a  large 
central  station  in  Chicago  gave  an  average  evaporation  of  6.1  pounds 
of  water  per  pound  of  coal,  actual  conditions,  whereas  the  evaporation 


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FINANCE  AND  ECONOMICS  —  COST  OF  POWER  703 

determined  by  dividing  the  total  water  fed  into  the  boiler  per  year 
by  the  total  consumption  of  coal  gave  only  5.2  pounds.  Current 
practice  gives  an  average  efficiency  (based  on  yearly  operation)  of 
boiler  and  furnace  of  70  per  cent  for  pumping  stations  running  at 
practically  full  load,  65  per  cent  for  large  lighting  and  power  stations 
with  yearly  load  factor  of  0.50  or  more,  and  60  per  cent  for  similar 
stations  with  load  factor  between  0.35  and  0.45.  For  very  low  load 
factors,  0.25  and  under  (as  in  connection  with  manufacturing  plants, 
tall  office  building  and  other  plants  operating  on  a  10-hour  basis) 
this  efficiency  seldom  exceeds  50  per  cent.  With  these  figures  as  a 
guide  the  cost  of  fuel  per  unit  output  may  be  roughly  approximated. 

400.  Oil,  Waste,  and  Supplies.  —  These  items  approximate  2  to  10 
per  cent  of  the  total  operating  expenses.     Tables  95  to  107  give  some 
idea  of  current  practice  in  different  classes  of  power  plants. 

401.  Repairs  and  Maintenance.  —  This  item  ordinarily  refers  to  the 
cost  of  keeping  the  plant  in  running  order  over  and  above  the  cost  of 
labor  or  attendance,  and  depends  upon  the  age  and  condition  of  the 
plant  and  the  efficiency  of  the  employees.     Tables  95  to  107  give  the 
cost  of   repairs    and    maintenance   for  a  wide   range  in  power-plant 
practice. 

402.  Cost  of  Power.  —  The  actual  cost  of  producing  power  depends 
upon  the  geographical  location  of  the  plant,  the  size  of  apparatus,  the 
design,  conditions  of  loading,  system  of  distribution,  and  the  method 
of  accounting.     Tables  95  to  108  compiled  from  various  sources  give 
the  detailed  costs  of  a  large  number  of  central  and  isolated  stations. 

Table  95.  Operating  costs  per  kilowatt  hour  for  a  number  of  typical 
British  electric  light  and  power  plants. 

Table  96.  Operating  costs  per  kilowatt  hour  for  a  number  of  United 
States  electric  power  plants  for  street-railway,  light,  and  general  power 
service. 

Table  97.  Average  operating  costs  per  kilowatt  hour  for  all  stations 
of  the  Boston  Elevated. 

Table  98.  Operating  costs  for  the  year  1907  of  the  mechanical  plant 
of  the  First  National  Bank  Building,  Chicago. 

Table  99.  Costs,  fixed  and  operating,  of  producing  one  brake  horse 
power  per  year,  simple  non-condensing  engine,  etc. 

Table  100.     Cost  of  power  for  compound  condensing  engine  plants. 

Table  101.  Costs  of  different  sizes  and  types  of  plants  and  annual 
costs  per  brake  horse  power,  average  working  conditions. 

Tables  102,  103,  and  104.  Cost,  fixed  and  operating,  of  producing 
one  electrical  horse  power  per  year  for  different  sizes  and  types  of 
plants. 


704 


STEAM  POWER  PLANT  ENGINEERING 


Table  105.  Influence  of  load  factor  on  cost  of  electrical  power  in 
isolated  stations. 

* 

Tables  106  and  107.  Cost  of  furnishing  heat,  light,  and  power  for  a 
number  of  isolated  stations  in  New  York  City,  tall  office  buildings,  loft 
buildings,  apartment  houses,  hotels,  and  club  buildings. 


Curves  showing  Range 
in  Cost  of  Power 
in  200  Mfg.  Plants 

Middle- Western  States 


I  1 


§811 

Size  of  Plant,  Horse-Power 
FIG.  418. 


Table  108.     Cost  of  furnishing  heat,  light,  and  power  for  the  year 
1907,  First  National  Bank  Building,  Chicago. 


FINANCE  AND  ECONOMICS  —  COST  OF  POWER 


705 


Reciprocating  steam- 


Steam-turbine  plant. 


Recipr|ocatingyengine  and  lo\\ 
turbine  plant 


G  as-en'gine  "plant. 
Gas-engine  and  steam 
6—  -Hydraulic-plant 


40  60 

PER  CENT  LOAD-FACTOR 

FIG.  418a.     Cost  of  Power  in  Large  Power  Plants  with  Maximum  Load  over 
30,000  Kilowatts.     Coal  at  $3.00  per  Ton.     14,500  B.T.U.  per  Pound. 


706 


STEAM  POWER  PLANT  ENGINEERING 


Dollars  per  Kilowatt  per  Year 

8  8  §  £  g  g 

1.=  Reciprocating  steam-plant. 
2.—  Steam-turbine  plant. 
3.  -Reciprocating-engine  and  low-pressure 
turbine  plant. 
4.  =  Gas-engine  plant. 
5.=  Gas-engine  and  steam-turbine  plant. 
6.  =  Hydraulic-plant 

V 

X 

X 

X 

x 

if 

>^ 

x 

x 

^ 

^ 

3A 

.X 

/\ 

^> 

^ 

^ 

X 

•^ 

% 

t^ 

^ 

^ 

^-^^ 

^ 

^ 

^ 

*:" 

^ 

20 


40  60 

Per  Cent  Load-Factor 


FIG.  418b.     Cost  of   Power  in  Large  Power  Plants  with  Maximum  Load  over  30,000 
kilowatts.     Coal  at  $1.50  per  Ton.     11,000  B.T.U.  per  Pound. 


FINANCE  AND  ECONOMICS  — COST  OF  POWER 


707 


Table  94a,  and  Figs.  418a,  418b  and  418c  give  the  fundamental 
relations  between  the  various  items  entering  into  the  cost  of  power  for 
various  types  of  plants  of  over  30,000  kilowatts  rated  capacity.  These 
data  are  taken  from  a  paper  presented  by  H.  G.  Stott  at  a  meeting  of 
the  Toronto  section  of  the  American  Institute  of  Electrical  Engineers, 
Toronto,  Ont.,  December,  1908.  The  figures  have  been  brought  up  to 
date  (June,  1910)  by  Mr.  Stott  and  show  what  is  actually  being  done 
to-day  in  large  plants  of  the  size  stated  above. 


TABLE  94a. 

DISTRIBUTION  OF  MAINTENANCE   AND    OPERATION    COSTS    IN    POWER  PLANTS 
HAVING  A    MAXIMUM  LOAD  OVER  30,000  KILOWATTS. 

(H.  G.  Stott.) 


Recip- 

Gas 

rocating 

T7*n 

Recii>- 

Steam 

Engines 

Gas 

Hin- 

rocating 

Tur- 

and Low- 

En- 

gines 
£Lnd 

Hy- 

dr3.ii- 

Steam 

bine 

pressure 

gine 

StGtini 

lir 

Plant. 

Plant. 

Steam 
Tur- 
bines. 

Plant. 

Tur- 
bines. 

1IC. 

Maintenance. 

1.    Engine  room,  mechanical  

2.59 

0.51 

1.55 

5.18 

2.84 

0.51 

2.   Boiler  or  producer  room  

*  4.65 

4.33 

3.55 

1.16 

1.97 

3.   Coal-  and  ash-handling  apparatus. 

0.58 

0.54 

0.44 

0.29 

0.29 

4.   Electrical  apparatus 

1.13 

1  13 

1   13 

1  13 

1  13 

1   13 

Operation. 

5.    Coal     . 

61.70 

55.53 

46.48 

26.52 

25.97 

6.   Water 

7.20 

0.65 

0.61 

3  60 

2.16 

7.   Engine  room,  labor  

6.75 

1.36 

4.06 

6^76 

4^06 

1.36 

8.   Boiler  or  producer  room,  labor  .  .  . 

7.20 

6.74 

5.50 

1.81 

3.05 

9.   Coal-  and  ash-handling,  labor  .... 

2.28 

2.13 

1.75 

1.14 

1.14 

...... 

10.   Ash  removal  

1.07 

0.95 

0.81 

0.54 

0.54 

1  1  .   Electrical  labor  

2.54 

2.54 

2.54 

2.54 

2.54 

2~54 

12.   Engine  room,  lubrication  

1.78 

0.35 

1.02 

1.80 

1.07 

0.20 

13.   Engine  room,  waste,  etc   

0.30 

0.30 

0.30 

0.30 

0.30 

0.20 

14.   Boiler  room,  lubrication,  etc  

0.17 

0.17 

0.17 

0.17 

0.17 

Relative  operating  cost,  per  cent  

100.00 

77.23 

69.91 

52.94 

47.23 

5.94 

Relative  investment,  per  cent  

100.00 

75.00 

80.00 

110.00 

96.20 

100.00 

Probable  average  cost  per  kilowatt  .  .  . 

125.00 

93.75 

100.00 

137.50 

120.00 

125.00 

Probable  fixed  charges  

11% 

11% 

11% 

12% 

11  5% 

11% 

/o 

/O 

/o 

/o 

A  A  t  V  /O 

•*•  •*•  /O 

For  steam-turbine  plants  larger  than  60,000  kw.  the  cost  per  kilowatt  may  be 

rliin^  to   *7S  00 


reduced  to  $75.00. 


708 


STEAM  POWER  PLANT  ENGINEERING 


Steam-turbinc-plantr-Cqst-$93r75 

00 


Hydraulic  plant.     Cost  $125.00  per  Kw. 


40 


60  80  100 

PERCENT  LOAD 


120 


140 


FIG.  418c. 


Cost  of  Power  in  Large  Power  Plant  with  Maximum  Load  over 
30,000  kilowatts. 


FINANCE  AND  ECONOMICS  —  COST  OF  POWER 


709 


TABLE  95. 

OPERATING  COSTS,  CENTS  PER  KILOWATT  HOUR.    TYPICAL  BRITISH  ELECTRIC 
LIGHT  AND   POWER  PLANTS,    1902. 

(Engineering  Record,  March  26,  1904,  p.  389.) 


Locality. 

Kilowatts 
Installed. 

Yearly 
Load 
Factor. 

Coal. 

Oil,  Waste, 
and 

Supplies. 

Wages. 

Re- 
pairs. 

Total. 

Bradford  

6,380 

Per  Cent. 
20.93 

0.52 

0.10 

0.16 

0.26 

1.04 

Leeds 

8  740 

12  31 

0  56 

0.06 

0.34 

0  28 

1  24 

St  Helen's 

1  340 

17  84 

0.52 

0.06 

0.34 

0  38 

1.3 

Edinburgh 

10  477 

14  75 

0.68 

0.08 

0.18 

0.36 

.3 

Bolton 

3  700 

18  87 

0  70 

0.12 

0.3 

0  20 

.32 

Booth            .    . 

850 

28  44 

0.82 

0.06 

0.3 

0.22 

.4 

Liverpool  
South  Shields 

21,190 
1,600 

25.11 
15  82 

0.74 
0.74 

0.12 
0.08 

0.3 
0.4 

0.26 
0  30 

.42 
.52 

Nottingham  

5,642 

12.97 

0.92 

0.20 

0.32 

0.18 

1.62 

Preston        

1,920 

13.31 

0.72 

0.12 

0.36 

0.46 

1.66 

Farnsworth  

610 

14.54 

0.92 

0.20 

0  36 

0.22 

1.7 

Leith  

990 

19.79 

1.10 

0.08 

0.42 

0.18 

1.78 

TABLE   96. 

OPERATING  COSTS,  CENTS  PER  KILOWATT  HOUR.     TYPICAL  STREET-RAILWAY 
PLANTS,   UNITED   STATES. 

(Street  Railway  Review,  Oct.  20,  1902,  p.  774.) 


Items. 

1 

2 

3 

4 

5 

Two 
Stations 
Supplying 
Power  for 
1000  Cars, 
175  Miles 
of  Track. 
(Belted  In- 
stallation.) 

One 
Station, 
Power  and 
Licht,   500 
Cars, 
230  ?Ciles 
of  Track. 

Large 
City 
Central 
Station. 

Combined  Light 
and  Power  Station. 

Inter, 
urban, 
Light 
and 
Power. 

July. 

November. 

Fuel        

0.68 
0.199 
0.013 
0.04 

0.005 
0.072 

0.008 
0.033 

0.803 
0.133 
0.018 

0.509 
0.115 
0.012 
.009 

0.958 
0.287 
0.048 

0.870 
0.276 
0.025 

1.004 
0.217 
0.023 
0.022 

0.015 
0.035 

0.010 

Wages 

Oil  and  waste  
Water  

Maintenance  : 
Buildings  

Steam  appliances  . 
Electric      equip- 
ment   
Miscellaneous  

Total  
Cost  of  coal,  dollars 
per  ton  
Output,  kilowatt 
hours  
Coal    per    kilowatt 
hour,  pounds  

0.029 

0.019 
0.023 

.011 

.001 
.012 

0.004 

1.05 
1.55 
20,981,295 
8.8 

1.03 
2.66 
2,140,641 
6.04 

0.669 
3.07 
641,650 

1.299 
2.98 
119,304 

6.4 

1.188 
2.98 
114,384 
5.55 

1.326 
3.62 
2,104,337 
5.55 

710 


STEAM  POWER  PLANT  ENGINEERING 


TABLE   97. 

OPERATING  EXPENSES  OF  BOSTON  ELEVATED  RAILWAY  COMPANY. 
AVERAGE  OF  ALL  STATIONS  AT  SWITCHBOARD,   APRIL   8,  1910. 


Rated  capacity,  kilowatts 

Yearly  load  factor 

Coal.*    Cents  per  K.W.H.. 
Labor.        "        "        " 
Repairs. t  "        "        " 
Supplies    "        "        "      .  . 

Total  "  "  .. 

Ratio  operating  expense 

to  gross  earnings 

Price  of  coal  per  ton 


Year. 


1902      1903      1904       1905         1906        1907       1908       1909 


0.58 


0.77 


35544  35544 


.391 
0.78 
0.18 

0.94 


.417 
0.53 
0.16 

0.67 


12    0.83 
0.6900.696 


35544 
.437 
0.45 
0.15 

0.57 
0.72 
0.68 


0.694 

$3. 60  $4. 854  $3. 55  $3. 1354  $3. 1859  $3. 572  $3. 568  $3 


38469 

.43 
0.47 
0.17 

0.60 


39969 

.45 
0.55 
0.18 

0.76 


0.94 
0.69 


50425 

.37 
0.56 
0.21 


1.07 
0.67 


50063 

.37 
0.45 
0.20 

0.61 


0.81 


0.654 
.209 


*  Coal  included  in  supplies. 


t  Repairs  included  in  supplies  and  labor. 


TABLE   98. 
COST  OF  OPERATION   (1907).     FIRST   NATIONAL   BANK  BUILDING,   CHICAGO. 


Coal  bill $34,567 . 27 

Ash  cartage 2,075.83 

Water  bill 3,535.  75 

Electrical  supplies  and  repairs 4,016 . 68 

Engine-room  supplies  and  repairs 1,311 . 36 

Boiler-room  supplies  and  repairs 1,316.50 

Oil,  waste,  and  grease 1 ,554 . 80 

Packings 727.01 

Machine-shop  supplies 187.13 

Refrigerating  supplies 761 . 81 

Steam  fitting 499.53 

Steam  heat  supplies 1 10.  20 

Plumbing  supplies 184 . 87 

Lamps 1,655.22 

Wages 24,572.89 

Petty  expenses 24 . 62 

Office  expenses 186 . 55 

Doctor  bills 43.00 

Coal  analysis , 160.00 

Total $77,591 . 02 

Total  receipts  from  all  sources  for  power,  heat,  light,  and  mis- 
cellany   71,435.85 

Net  cost.  .  $6,055.17 


FINANCE  AND  ECONOMICS  —  COST  OF  POWER 


711 


TABLE   99. 

COST   OF  ONE  HORSE  POWER  PER  YEAR,   SIMPLE  ENGINES,  NON-CONDENSING, 
10-HOUR  BASIS,   308  DAYS  PER  YEAR. 

(Wm.  O.  Webber,  Engineer  U.S.,  March  16,  1903,  p.  241.) 


Size  of  plant  horse  power 
Cost  of  plant  per  horse  power  
Fixed  charges  at  14  per  cent  

20 
$200.00 
28.00 

40 
$190.00 
26.60 

60 
$180.00 
25.20 

80 
$175.00 
24.50 

Coal  per  horse-power  hour,  in  pounds.  .  .  . 
Cost  at  $4  00  per  ton 

12.00 
66  00 

10.00 
55  00 

9.00 
49  50 

8.00 
44  00 

Attendance  10  hour  basis  

30.00 

20.00 

15.00 

13.00 

Oil  waste  and  supplies 

6  00 

4  00 

3  00 

2  60 

With  coal  at  $5  00  per  ton 

146  50 

119  35 

105.07 

95  10 

With  coal  at  $4  50  per  ton 

138  25 

112.47 

98  80 

89  60 

With  coal  at  $4.00  per  ton  
With  coal  at  $3.50  per  ton  
With  coal  at  $3.00  per  ton  :  
With  coal  at  $2.50  per  ton  
With  coal  at  $2  00  per  ton  

130.00 
121.75 
113.50 
105.25 
97.00 

105.60 
98.72 
91.85 
84.97 
78.10 

92.70 
86.51 
80.32 
74.13 
67.95 

84.10 
78.60 
73.10 
67.60 
62.10 

TABLE    100. 

COST  OF  ONE  HORSE  POWER  PER  YEAR,  COMPOUND  CONDENSING  ENGINES, 
10-HOUR  BASIS,   308  DAYS  PER  YEAR. 

(Wm.  O.  Webber,  Engineer  U.S.,  Feb.  2,  1903,  p.  144.) 


Size  of  plant                  horse  power 

100 

200 

300 

400 

500 

600 

Cost  of  plant  per  horse  power.  .  .  . 

$170.00 

$146.00 

$126.00 

$110.00 

$96.00 

$85.00 

Fixed  charges  at  14  per  cent  

23.80 

24.40 

17.65 

15.40 

13.45 

11.90 

Coal  per  horse-power  hour,  pounds 

7.0 

6.5 

6.0 

5.5 

5.0 

4.5 

Cost  of  fuel  at  $4.00  per  ton  

38.50 

35.70 

33.00 

32.00 

27.50 

24.70 

Attendance,  10-hour  basis  

12.00 

10.00 

8.60 

7.25 

6.20 

5.40 

Oil,  waste,  supplies  

2.40 

2.00 

1.72 

1.45 

1.24 

1.08 

Total 

76.70 

68.10 

60.97 

56.10 

48.39 

43.08 

With  coal  at  $5.00  per  ton  

86.40 

77.10 

69.22 

61.90 

55.29 

49.28 

With  coal  at  $4.50  per  ton  

81.50 

72.60 

65.07 

58.10 

51.79 

46.18 

With  coal  at  $4.00  per  ton  

76.70 

68.10 

60.97 

56.10 

48.39 

43.08 

With  coal  at  $3.50  per  ton  

71.90 

63.70 

56.82 

50.50 

45.04 

30.98 

With  coal  at  $3.00  per  ton  

67.00 

59.20 

51.67 

46.70 

41.49 

36.88 

With  coal  at  $2.50  per  ton  

62.30 

54.75 

48.59 

43.00 

38.83 

33.83 

With  coal  at  $2  00  per  ton 

57.45 

50.25 

44.47 

40.10 

34.64 

30.73 

Size  of  plant  horse  power 

700 

800 

900 

1000 

1500 

2000 

Cost  of  plant  per  horse  power.  .  .  . 

$76.00 

$69.00 

$64.00 

$60.00 

$58.00 

$56.00 

Fixed  charges  at  14  per  cent  

10.65 

9.65 

8.95 

8.40 

8.12 

7.85 

Coal  per  horse-power  hour,  pounds 

4.0 

3.5 

3.0 

2.5 

2.0 

1.5 

Cost  of  fuel  at  $4  00  per  ton 

22.00 

19.20 

16.50 

13.75 

11  .00 

8.25 

Attendance    10-hour  basis 

4.70 

4.15 

3.75 

3.50 

3.25 

3.00 

Oil  waste  supplies 

0.94 

0.83 

0^75 

0.70 

0.65 

0  60 

Total 

38.29 

33^83 

29^95 

26^35 

23  '02 

19^70 

With  coal  at  $5.00  per  ton  

43.79 

39.73 

34.05 

29.80 

25.77 

21.75 

With  coal  at  $4.50  per  ton  

41.04 

36.28 

32.00 

28.05 

24.39 

20.72 

With  coal  at  $4.00  per  ton  

38.29 

33.83 

29.95 

26.35 

23.02 

19.70 

With  coal  at  $3.50  per  ton  

35.54 

31.48 

27.87 

24.60 

21.64 

18.67 

With  coal  at  $3.00  per  ton  

32.79 

29.03 

25.80 

22.90 

20.27 

17.65 

With  coal  at  $2.50  per  ton  

30.04 

27.18 

23  75 

21  20 

18.89 

16.60 

With  coal  at  $2.00  per  ton  

27.29 

24.23 

21.70 

19.47 

17.52 

15.57 

712 


STEAM  POWER   PLANT  ENGINEERING 


TABLE   101. 

SHOWING  CAPITAL  COSTS  OF  PLANTS  INSTALLED  AND  ANNUAL  COSTS  OF  POWER 
PER  BRAKE  HORSE  POWER,  AVERAGE  WORKING  CONDITIONS. 

(H.Von  Schon,  Engineering  Magazine,  May,  1907.) 

CLASS  I.  —  Engines:   Simple,  Slide- Valve, Non-Condensing. 
Boilers:   Return  Tubular. 


Size  of  Plant, 
Horse  Power. 

Engines, 
Boilers,  etc., 
Installed. 

Capital  Cost  of  Plant  per 
Horse  Power  Installed. 

Annual  Cost 
of  10  Hours 
Power  per 
B.H.P. 

Annual  Cost 
of  20  Hours 
Power  per 
B.H.P. 

Building. 

.Total. 

10  

$66.00 
56.00 
48.70 
44.75 
43.00 

$40.00 
37.00 
35.00 
33.50 
31.00 

$106.00 
93.00 
83.70 

78.25 
74.00 

$91.16 
76.31 
66.46 
59.49 
53.95 

$180.76 
151.48 
131.68 
117.74 
106.46 

20  

30 

40 

50  

CLASS  II.  —  Engines:   Simple,  Corliss,  Non-Condensing. 
Boilers:   Return    Tubular. 


30  

$70.70 

$35.00 

$105.70 

$61.14 

$117.70 

40 

62  85 

33  50 

96.35 

55.50 

107  10 

50 

59  00 

31.00 

90.00 

50.70 

97  73 

60     

56.00 

30.00 

86.70 

47.42 

91.34 

80   

50.00 

27.50 

77.50 

43.86 

85.41 

100  

44.60 

25.00 

69.60 

40.55 

79.19 

CLASS  III.  —  Engines:  Compound,  Corliss,  Condensing: 
Boilers:  Return  Tubular  with  Reserve  Capacity. 


100  
150  
200 

$63.40 
53.70 
50  10 

$28.00 
24.00 
20.00 

$91.40 
77.70 
70.10 

$33.18 
29.83 
28.14 

$60.05 
54.63 
51.72 

300 

45  90 

18.00 

63.90 

26.27 

48.83 

400  
500  

43.55 
41.25 

16.00 
14.00 

59.55 
55.25 

24.84 
23.73 

46.12 
44.21 

750  
1000 

40.50 
39  00 

13.00 
12.00 

53.50 
51.00 

23.56 
23.26 

44.02 
43.71 

FINANCE  AND  ECONOMICS  —  COST  OF  POWER 


713 


TABLE   102. 

COST  OF  ELECTRICAL  POWER  (W.  M.  WILSON). 
(Power,  January,  1907,  p.  15.) 


CAPACITY  Or  'J 
W.ANTI.H.P.  1 

ii 

»*> 

(L 

Z 

1 

NO.OF  ENGINES! 

il 

COST  OF 

GENERATORS 

•1 

8| 

sl 

it 

I 

TYPE  OF 
BOILERS 

s! 

BOILERS 

II 

0  0 

o  « 

>! 

§| 

COST  OF 

ECONOMIZERS  | 

COST  OF 
HEATERS 

•oSKcSfr, 

H.B, 

00 

30 

00 

2 

2880 

2300 

1000 

1292 

92 

F.T. 

2 

2 

1904 

2680 

280 

160 

0 

COMPOUND 
NON-CON*. 
H.&. 

20 

24 

00 

2 

4600 

2300 

1000 

1430 

74 

.T. 

2 

4 

1640 

2510 

230 

130 

O 

COMPOUND 
CONO.  H.  8. 

20 

20 

100 

2 

4600 

2300 

1000 

1400 

1430 

66 

.T. 

2 

66 

1500 

2414 

200 

125 

DE  I  AVAL 

TURBINE 

20 

8.6 

100 

2 

7400 

1000 

1200 

910 

62 

.1. 

2 

62 

1425 

2360 

200 

120 

SIMPLE 
HON-COhl). 
M.S. 

00 

30 

100 

3 

4290 

3450 

1500 

1830 

185 

F.T. 

3 

93 

2650 

3850 

630 

235 

s 

COMPOUND 
NON-CONO. 
H.B. 

20 

24 

100 

3 

G300 

3450 

1500 

2040 

146 

.T. 

3 

74 

2460 

3600 

440 

200 

CM 

COMPOUND 
COND.  M.  8. 

20 

70 

100 

3 

6900 

8450 

1500 

2C50 

2040 

132 

F.T. 

3 

66 

2250 

3440 

400 

180 

DE  LAVAL 
TURBINE 

20 

8.6 

100 

3 

11100 

1500 

1940 

1270 

24 

.T. 

2 

24 

2480 

2825 

390 

180 

SIMPLE 
NON-COND. 
H.B. 

00 

30 

200 

3 

7020 

6630 

3000 

2450 

370 

F.T 

4 

23 

4920 

5312 

1000 

400 

COMPOUND 
NON-COND. 
H.  6. 

20 

24 

200 

3 

10620 

6630 

3000 

2698 

297 

F.T. 

4 

99 

3960 

5098 

840 

335, 

0 

COMPOUND 
COND.  H.  8. 

20 

20 

200 

3 

10620 

6630 

3000 

3700 

2700 

264 

F.T. 

4 

88 

3670 

4970 

740 

312 

* 

OE  LAVAL 
TURBINE 

20 

17 

200 

3 

18960 

3000 

3250 

1620 

22c 

F.T. 

3 

13 

3390 

4000 

650 

275 

VERTICAL 
COND.  L.  S. 

50 

13 

400 

2 

20400 

10600 

4000 

2600 

4300 

153 

W.T 

2 

153 

6100 

2300 

500 

C65 

200 

HOR. 
COND.  L.  8. 

50 

13 

400 

2 

13600 

15400 

4000 

2600 

5440 

153 

W.T. 

2 

153 

6100 

2300 

500 

6C5 

200 

§ 

to 

SIMPLE 

NON-CONO. 
M.S. 

100 

30 

200 

4 

9360 

8840 

4000 

3190 

555 

FT. 

6 

111 

6660 

.  7070 

1460 

560 

COMPOUND 
NON-COND. 

M.S. 

120 

24 

230 

4 

141CO 

8840 

4000 

3512 

445 

F.T. 

5 

111 

5550 

6440 

1200 

400 

120 

20 

200 

4 

14160 

8840 

4000 

6050 

3512 

397 

F.T 

5 

99 

4950 

6296 

1070 

420 

COND  M.8. 

OE  LAVAL 

T20 

17 

2uO 

4 

25280 

4000 

4500 

2094 

33S 

F.T. 

4 

113 

4520 

5240 

930 

380 

VERTICAL 
COND.  L.  B. 

150 

13 

300 

3 

29550 

12600 

4500 

3600 

62J5 

230 

W.I 

2 

230 

8460 

2530 

650 

1150 

280 

HOR. 
COND.  I.  8. 

150 

13 

300 

3 

19800 

19920 

4500 

360U 

76CO 

i30 

W.T 

2 

230 

8460 

2530 

650 

1150 

260 

PARbOKS 

TURPINF 

150 

13 

000 

2 

35010 

bOOO 

3600 

1470 

230 

W.T 

2 

230 

8460 

2530 

650 

1150 

280 

§ 

CM 

NON-CONO 
M.S. 

100 

30 

400 

4 

15200 

18000 

8000 

4030 

111 

F.T 

10 

12C 

12300 

12778 

2500 

1060 

COMPOUN 
NON-COND 

H.  S. 

120 

24 

400 

4 

22280 

18000 

8000 

4540 

89C 

F.T 

8 

127 

10160 

10340 

2120 

UCO 

COND.  H.  8 

120 

20 

400 

4 

22280 

18000 

8000 

8100 

4540 

794 

F.T 

8 

113 

9040 

10128 

1950 

780 

DE  LAVAL 

TURBINE 

120 

166 

300 

5 

45030 

7500 

7200 

2960 

660 

F.T. 

7 

110 

7700 

8880 

1700 

6GO 

COND.  L.  8 

150 

13 

600 

3 

43200 

21300 

9000 

6100 

7132 

460 

W.T 

3 

230 

12690 

3660 

1220 

2300 

480 

COND.  L.  8 

150 

13 

000 

3 

28800 

28020 

9000 

6100 

8420 

400 

W.T 

3 

23u 

12690 

3660 

1220 

2300 

480 

PARSONS 

150 

13 

1200 

2 

45000 

12000 

6100 

2800 

460 

W.T 

3 

230 

12690 

3660 

1220 

2300 

480 

o 

§ 

COND.  1.  8 

150 

13 

500 

5 

60000 

31350 

12500 

8450 

10735 

76C 

W.T 

4 

250 

1C400 

4900 

1£20 

3840 

760 

CONO.  I.  8 

150 

13 

1000 

3 

48000 

34650 

15000 

•8450 

9320 

76 

W  T 

4 

256 

18400 

4900 

1920 

3340 

700 

TURBINE 

150 

13 

100 

3 

62400 

15000 

8450 

3360 

76 

W.T 

4 

256 

18400 

4EOO 

1620 

384 

700 

714 


STEAM   POWER   PLANT   ENGINEERING 


TABLE   102   (Continued). 


ll 

H    I 
*"  I 

M                                               CONTINUOUS    EXPENSE   PER   YEAR 

INTEREST 
DEPRECIATION 
REPAIR?  -TAX 
INSURANCE 

Is 

a 

I 

(9 

COAL                                   g                                       TOTAL 

CONOENSIN 
WATER 

UPPER  LINE      10  Hrt.    DAY 
LOWER  LINE     24  HR.   DAY 

COST  OF  COAL  PER  2000  LB 

COST  OF  COAL  PER  2000 

LB 

$700 

S5.00 

$3.00 

SI.  50 

$7.00 

$5.00 

$300 

Si  .40 

4GO 

12956 

1814 

102 

1550 

4474 

3197 

1913 

959 

7940 

6663 

6384 

4426 

289 

4380 

11234 

8024 

4814 

2407 

17717 

14507 

1129" 

8 

UU 

.296 

14136 

102 

1550 

3594 

2568 

1540 

770 

7225 

6198 

517 

V 

£i 

289 

4380- 

9022 

6444 

3867 

1933 

16670 

13092 

10514 

• 

81 

264 

15233 

102 

1650 

193 

3206 

2291 

13/4 

687 

7184 

6269 

5352 

4665 

289 

4380 

646 

8049 

6749 

3449 

1725 

15397 

13097 

10797 

90/3 

248 

14863 

102 

1550 

180 

2986 

2133 

1280 

640 

6899 

6046 

6192 

4553 

289 

4380 

508 

7478 

6348 

3209 

1604 

14736 

I26O6 

1046' 

1 

8862 

740 

19275 

2699 

205 

21/U 

8949 

6392 

3836 

1U1S 

14023 

11466 

8910 

6992 

678 

6132 

22468 

16048 

9628 

4814 

31877 

25457 

19037 

14223 

592 

21182 

205 

21/0 

7188 

5135 

3081 

1540 

12528 

10475 

842 

6880 

6/8 

6132 

18044 

12888 

7/33 

3867 

27/19 

22563 

17408 

135 

42 

623 

22738 

3183 

205 

217U 

387 

6413 

4581 

2/48 

1374 

.12358 

10526 

8693 

7319 

678 

6132 

1093 

16097 

11498 

6899 

3449 

2/083 

22484 

17886 

14435 

496 

22181 

3106 

205 

2170 

360 

5972 

4266 

2559 

128U 

11812 

IOIO6 

839S 

7120 

5/tt 

6132 

1016 

14956 

10696 

6418 

8209 

25787 

21527 

I724S 

14U40 

1480 

32212 

4510 

409* 

8026 

17898 

12784 

7671 

3836 

25843 

20729 

15616 

11781 

1166 

8550 

44936 

32096 

19266 

9628 

59162 

46312 

33472 

23844 

1188 

34369 

4812 

409 

3026 

14377 

10269 

6161 

3081 

22624 

160IC 

14408 

11328 

1156 

8550 

86088 

25777 

16466 

7733 

50606 

40295 

29984 

22251 

1056 

37398 

6236 

409 

3026 

773 

12625 

9101 

6497 

2748 

22269 

18605 

14941 

12192 

1156 

866U 

2187 

321S5 

22996 

13798 

6399 

49324 

40125 

80927 

24028 

904 

36049 

60<7 

40S 

8026 

658 

10913 

7795 

46// 

2339 

20053 

16935 

13817 

114/tf 

1160 

8660 

1858 

27373 

19552 

11/31 

6866 

43984 

36103 

28342 

2247/ 

612 

52277 

7318 

409 

3026 

"603" 

7420 

6300 

3180 

1690 

18676 

16556 

14436 

12847 

1156 

8550 

1421 

18615 

13296 

7978 

3988 

37060 

31741 

26423 

22433 

612 

61417 

7i98 

409 

3026 

603 

7420 

6300 

81MU 

1590 

18556 

16436 

V4316 

12727 

1166 

856* 

1421 

,    18615 

13296 

7978 

tftftttt 

36940 

31621 

263O3 

22313- 

2220 

43960 

6153 

614 

8813 

26848 

19177 

11606 

5763 

37418 

29747 

22078 

16323 

1734 

107/5 

67404 

48144 

28884 

14442 

86056 

66796 

4753G 

33094 

'1780 

45942 

6432 

614 

3813 

21565 

15404 

92*2 

4621 

32424 

26263 

20100 

I548O 

1734 

10/75 

54132 

38665 

23199 

uouo 

73073 

67006 

42140 

30541 

1588 

49886 

6984 

614 

3813 

1161 

19238 

13733 

8245 

4122 

31810 

26305 

20817 

16894 

1734 

10775 

32/9 

48292 

84494 

20697 

10348 

71064 

67266 

.  43469 

33120 

1356 

48300 

67d2 

614 

8813 

986 

16370 

11693 

7016 

8608 

28545 

2386b 

19191 

16C83 

1734 

10775 

2783 

41059 

29328 

17597 

8799 

63118 

51387 

89050 

80850 

920 

69465 

9725 

614 

3813 

755 

11130 

7950 

4770 

2386 

26037 

22867 

19677 

17293 

1734 

10775 

21132 

27923 

19945 

11967 

6983 

52289 

44311 

36833 

30349 

920 

69490 

9729 

014 

3C13 

755 

11130 

7950 

4770. 

2386 

26041 

22861 

19681 

17297 

1734 

10775 

2132 

27923 

1934a 

11967 

5983 

62293 

44315 

86337 

30353 

920 

60060 

8408 

614 

3813 

755 

1113U 

7950 

4770 

2386 

2472O 

2I54O 

I836O 

16976 

1734 

10776 

2132 

27923 

19945 

11967 

6983 

50972 

42994 

35016 

29032 

4440 

78806 

10963 

1227 

6680 

53694 

38353 

23012 

11506 

71464 

66123 

40782 

29276 

34t>H 

16-768 

134BWJ 

96286 

67708 

28884 

165008 

126488 

87908 

6901 

4 

3560 

79860 

11180 

\i'J.t 

6580 

43I2U 

30807 

lb4bo 

9242 

61116 

46/94 

36472 

27228 

3469 

16768 

108263 

77331 

•  46398 

23200 

138680 

107/48 

76815 

63617     { 

3176 

86994 

12039 

1227 

6580 

232  1 

38476 

274C5 

104Stf 

8246 

59643 

48632 

3/656 

29412 

3469 

15768 

6558 

965o4 

68989 

413U3 

20697 

134418 

106823 

79227 

68531 

2640 

84240 

11794 

1277 

6580 

1925 

31965 

29830 

136«y 

6850 

5249) 

43356 

34226 

27376 

3469 

15768 

5440 

80207 

57290 

34374 

17187 

116678 

93/61 

70845 

63658 

1840 

108992 

15249 

1227 

6580 

1509 

22260 

15900 

9640 

4772 

45825 

39465 

"  331U6 

28337 

34o« 

16768 

4264 

65846 

89890 

23934 

11969 

94596 

78040 

62684 

5071 

1840 

102530 

14354 

6580 

1509 

2226U 

15900 

9540 

4772 

44930 

36570 

,  32210 

27442 

3469 

15768 

4264 

55840 

39890 

28934 

11969 

93701 

77745 

61/89 

49S24 

1840 

88090 

12333 

mi 

5580 

1509 

22260 

15900 

954U 

4772 

429O9 

36549 

30189 

25421 

3469 

15768 

4264 

65846 

39890 

23934 

11969 

91680 

71724 

59768 

4780J 

* 

3072 

155927 

21830 

2046 

7440 

2615 

15901 

7952 

70«Jl 

60331 

49/32 

41783 

5782 

21025 

7106 

93076 

66483 

39890 

1994 

148818 

122225 

95632 

75685 

3072 

141312 

20764 

2046 

7440 

2516 

3710U 

26500 

lObU 

7«52 

69865 

59265 

48666 

40717 

5/82 

21025 

7106 

93076 

66483 

39890 

19943 

14/752 

121159 

945GS 

74619 

3072 

122102 

17094 

204b 

7440 

2516 

87100 

26500 

15901 

/952 

66195 

55595 

44996 

370*7 

6782 

21025 

7103 

93076 

66483 

39890 

19943 

I44O82 

1  1  7489 

9O896 

7O949 

FINANCE  AND  ECONOMICS  -  COST  OF  POWER 


715 


TABLE   103. 

COST   OF   ONE  HORSE   POWER   PER  YEAR  IN  STREET-RAILROAD  SERVICE  FOR 
DIFFERENT  CLASSES  OF  ENGINES. 

1000-Horse-Power  Plant.     (R.  C.  Carpenter.) 
(Sibley,  Journal  of  Engineering,  November,  1904,  p.  92.) 


. 

Tons  per  Horse 

00   0 

tS 

^ 

1 

Power  per  Year. 

^-S 

J 

-  & 

^ 

f? 

K 

73 

§2 

1    . 

1 

i 

| 

o 

o 

9 

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i 

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tj 

W 

M 

W 
oo 

n 

1  8,d 

1^3  **g 

li 

ftf 

1 

7 

1 

8  e  H 

I 

1  1^' 

-  1 

"3  S 

| 

| 

"«   o  fc 

j 

^  g  § 

"i  s 

3 

Q 

Q 

Q 

£ 

0 

^ 

H 

Non-Condensing  Engines. 

Simple  slide-valve,    aver- 

Q  O"P 

4  63 

10.  14 

15.21 

20.28 

on  4.0 

7.30 

4  44 

42   1fi 

Simple  slide-valve,  best.  .  . 
Simple  Corliss,  average  

4^60 
3.45 

7^55 

15.10 
11.33 

20.14 
15.10 

Ol/  .  T:*. 

30.20 
22.66 

7^30 
7.30 

Tt  .  rtrt 

4.44 
4.75 

^t^«  JlU 

41.94 
34.71 

Simple  Corliss,  best  

3.01 

6.59 

9.89 

13.18 

19.78 

7.30 

4.75 

31.83 

*Compound  slide-valve.  .  .  . 

4.17 

9.05 

13.57 

18.10 

27.14 

6.90 

4.80 

38.84 

Condensing  Engines. 

Compound         slide-valve, 

average  

3.25 

7.12 

10.68 

14.24 

21.36 

6.70 

4.72 

32.78 

Compound          slide-valve, 

best  

2.40 

5.25 

7.88 

10.51 

15.76 

6.50 

4.72 

26.98 

Compound  Corliss,  average 
Compound  Corliss,  best.  .  . 

2.36 
1.80 

5.17 
3.94 

7.74 
5.91 

10.33 

7.88 

15.48 
11.82 

6.50 
6.10 

5.28 
5.28 

27.26 
23.20 

*  The  compound  slide-valve  engine,  running  non-condensing,  made,  in  this  series  of  tests,  a 
poorer  record  than  the  single  Corliss.  This  may  have  been  due  to  the  extremely  bad  conditions 
of  loading.  It  is,  I  think,  a  fact  that  this  class  of  engine  has  not  been  a  marked  success  for 
street  railway  work. 


716 


STEAM  POWER  PLANT  ENGINEERING 


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FINANCE  AND  ECONOMICS  —  COST  OF  POWER 


717 


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718 


STEAM  POWER  PLANT  ENGINEERING 


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FINANCE  AND  ECONOMICS  —  COST  OF  POWER 


719 


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720 


STEAM  POWER   PLANT  ENGINEERING 


TABLE   105. 

COST  OF  POWER. 
Examples  of  Isolated  Station  Practice. 


Manufactur- 

Large Office 

Small  Office 

ing  Plant, 

Building. 

Building. 

Electro- 

plating. 

Rated  capacity     .    .  . 

kilowatts 

500 

50 

500 

Yearly  capacity  

.  .  .  kilowatt  hours 

4,380,000 

438,000 

4,380,000 

Actual  load      

.  kilowatt  hours 

670,000 

40,820 

3,500,000 

Yearly  load  factor.  .  .  . 

per  cent 

17 

9.1 

'  80 

Curve  load  factor  

per  cent 

17 

24.7 

80 

Operating  Charges,  per  Year. 


Labor  

$6,050  00 

$1  400  00 

$12  300  00 

Coal  and  ashes  

6,342.00 

960  00 

9  100  00 

Water 

642  00 

75  00 

Oil  and  waste 

168  00 

90  00 

210  00 

Lamps            .       .         ... 

395  00 

41  00 

50  00 

Repairs  and  renewals       

69  00 

182  00 

1  008  00 

Total 

$13  666  00 

$2  748  00 

$22  668  00 

Fixed  Charges,  per  Year. 


Interest  (5  per  cent) 

$3  500  00 

$325  00 

$4  500  00 

Depreciation  (6  per  cent) 

4  200  00 

628  00 

5  400  00 

Insurance  (J  per  cent)              

350  00 

30  00 

450  00 

Taxes  (1|  per  cent)       

1,050.00 

90  00 

1  350  00 

Rental          

900  .  00 

Total  

$10,000.00 

$1,073.00 

$11,700.00 

Cost  per  Kilowatt  Hour,  Cents. 


Operating  charges          

1.79 
1.32 
3.11 

6.78 
2.62 
9.40 

.65 
.33 

.98 

Fixed  charges     

Total  cost   

FINANCE  AND  ECONOMICS  —  COST  OF  POWER 


721 


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722 


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FINANCE  AND  ECONOMICS  — COST  OF  POWER 


723 


TABLE   107a. 

YEARLY   OPERATING  COSTS  IN   FOUR  TYPICAL  CENTRAL  STATIONS, 

STATE   OF  MASSACHUSETTS 

Year  ending  June,  1909. 


Salem  Elec- 
tric Light 

Fitchburg  Gas 
&  Electric 

Haverhill 
Electric  Co. 

Maiden 
Electric  Co. 

Co. 

Co. 

Type  of  Prime  Mover 

6  Engines 

3  Engines 

(  2  Turbines 
{    1  Engine 

(  1  Turbine 
(  3  Engines 

Rated  station  capacity   kw.. 

2500 

2000 

2300 

Output,  millions  of  kw.  hrs..  . 

3.106 

4.006 

3.721 

4.715 

Yearly  load  factor,  per  cent. 

14.2 

22.9 

18.5 

Total  station  operating  force 

14 

12 

13 

14 

Cost  of  f  uel  ,  dollars  per  ton  .  . 

4.51 

4.52 

3.97 

3.78 

Coal  per  kw.  hr  

3.3 

3.28 

3.27 

3.02 

Operating  Costs,  Cents  per  Kilowatt  Hour. 


Coal 

0  740 

0.740 

0.650 

0.565 

Oil  and   waste                     .    .  . 

0.025 

0.015 

0.190 

0.020 

Water                        

0.027 

0.025 

0.003 

0.045 

Wages 

0.410 

0.308 

0.285 

0.320 

Station  building  repairs  
Steam  equipment  repairs.  .  .  . 
Electrical  equipment  repairs. 
Miscellaneous 

0.034 
0.158 
0.011 

0.017 
0.041 
0.072 
0  024 

0.063 
0.073 
0.019 
0  040 

0.023 
0.072 
0.14 
0  21 

Total                      

1  412 

1.242 

1.152 

1  08 

TABLE    107b. 

COST  OF  POWER,  CENTS  PER  KW.  HOUR.      STEAM-ELECTRIC  CENTRAL  STATIONS. 

Year  ending  June  30,  1908. 


Bos- 
ton. 

Worcester. 

Lowell. 

Fall 
River. 

Mai- 
den. 

Cam- 
bridge. 

Lynn. 

Fuel 

462 

703 

710 

880 

635 

690 

618 

Oil  and  waste 

008 

027 

009 

032 

017 

019 

012 

Water  

024 

.034 

008 

012 

032 

055 

040 

Wages  

.192 

.360 

.262 

538 

342 

347 

296 

Station  repairs  

.015 

.012 

020 

012 

035 

021 

052 

Steam  repairs  

.042 

.055 

020 

037 

072 

059 

147 

Electrical  repairs 

056 

055 

009 

029 

014 

046 

045 

Miscellaneous 

023 

000 

022 

080 

033 

000 

000 

Total  

.822 

1.246 

1  060 

1  620 

1  180 

1  237 

1  210 

Cost  of  fuel  per  ton  
Output,  millions    kilowatt 
hours  per  year  
Capacity  of  station,  thou- 
sands of  H.P  

3.99 
88.5 
73.5 

4.79 
5.4 
5.90 

4.75 
9.4 
7.39 

4.68 
4.0 
4.43 

4.49 
4.6 
4.87 

4.40 
6.0 
6.75 

3.60 

8.7 
8.2 

724  STEAM  POWER  PLANT  ENGINEERING 

TABLE   108. 

COST    OF    POWER    (1907),  FIRST    NATIONAL   BANK   BUILDING,    CHICAGO. 


Total  weight  of  coal  burned tons  14,956 

Total  weight  of  water  evaporated gallons  22,100,000 

Total  electrical  output kilowatt  hours  1,546,600 

Water  actually  evaporated  per  pound  of  coal  as  fired 6.1 

Electric-Light  Plant.     (Cost  of  Power  only.) 

Pounds  of  coal  per  kilowatt  hour 6.57 

Cost  of  coal  per  kilowatt  hour cents  0 . 88 

Cost  of  labor  per  kilowatt  hour cents  0 . 78 

Cost  of  supplies  per  kilowatt  hour cents  0.11 

Total 1 . 77 

All  expenses  of  entire  plant  charged  against  switchboard. 

Pounds  of  coal  per  kilowatt  hour 19 . 34 

Cost  of  coal  per  kilowatt  hour cents  2 . 23 

Cost  of  labor  per  kilowatt  hour. cents  1 .59 

Cost  of  supplies  per  kilowatt  hour cents  1 . 18 

Total 5.00 

Elevator  Plant. 

Passengers  handled 2,016,300 

Total  car  miles 92,700 

Cost  of  labor  per  car  mile cents  2 . 02 

Cost  of  material  per  car  mile    cents  3 . 12 

Cost  of  power*  per  car  mile  cents  3.00 

Total..  8.14 


*  Approximate. 

BIBLIOGRAPHY.     COST  OF  POWER.  * 
ELECTRICAL. 

Rented   Power  for   Electric   Railways,    American 

Electrician 10:  329  July,  1898 

Cost  of  Pumping  Station  in  New  York,  Electrical 

World 43:  820  April  30,  1904 

Cost  of  Electric  Power  in  Comparison  with  Steam 

for  Traction,  Engineer  (London) 90:  600  Dec.  14,  1900 

Cost  of  Steam  and  Electricity,  Engineering 74:  667  Nov.  21,  1902 

Cost  of  Power,  Engineering 76:  706  Nov.  20,  1903 

Economy    of    Isolated    Plant    (I.    D.    Parsons), 

Engineering  Magazine 22:  573  Jan.,  1902 

720  Feb.,  1902 


*  See  page  729  for  bibliography  1908-1910. 


FINANCE  AND  ECONOMICS  —  COST  OF  POWER 


725 


ELECTRICAL  —  Continued. 

Cost  of  Energy  in  Electric  Supply  (A.  D.  Adams), 

Engineering  Magazine  

24: 

181 

Nov.,  1902 

Street  Railway  Review  

12: 

149 

March,  1902 

Data  on  Electric  Power  Generation  in  Glasgow, 

Engineering  Record  

478 

April  22,  1899 

Cost   of   Generating  Electric   Power    (E.  J.  Fox), 

Engineering  Record  

49: 

388 

March  26,  1904 

Relative  Cost  of  Electric  Power  for  Three  Types  of 

Plants  (R.  D.  Mushon),  Engineering  Record  

49: 

411 

April  2,  1904 

Cost  of  Electric  Power  for  Street  Railways  (R.  W. 

Conant),  Power  

18: 

8 

Oct.,  1898 

Street  Railway  Review  

8: 

631 

Sept.,  1898 

Cost  of  Power  in  New  Orleans  Railroad  Company 

Power  Plant,  Street  Railway  Journal  

18: 

668 

Nov.  2,  1901 

Cost    of     Power,    Union    Traction    Company    of 

Indiana,  Street  Railway  Journal  

18: 

827 

Dec.  7,  1901 

Relative  Costs  of  Steam  and  Polyphase  Traction, 

Street  Railway  Journal  

21: 

737 

May  16,  1903 

Cost  of  Power  at  Newcastle-on-Tyne,  Street  Rail- 

way Journal  

22: 

207 

Aug.  8,  1903 

Charges  for  Rented  Power,  Street  Railway  Review.  .  . 

8: 

236 

April,  1898 

Cost  of  Power  for  Electric  Railways,  Street  Railway 

Review  

g: 

43  97,  186,  224,  340,  385, 

461 

,  760,  886,  1898 

9: 

35, 

123,   185,  261,   319, 

459 

,  529,  595,  749,  851 

10: 

11,  223,  399,  521,  735 

11: 

123, 

416,  418,  1901 

Cost  of  Niagara  Power  at  Buffalo,  Street  Railway 

Review  

8: 

339 

May,  1898 

Statistics  on  the  Cost  of  Power,   Street  Railway 

Review  

12: 

77 

Oct.,  1902 

Analysis  of  Cost  of  Generation  and  Distribution 

of  a  Unit  of  Electricity  (C.  W.  Rice),  Western 

Electrician  

22: 

574 

June  25,  1898 

Cost  of  Electric  Power  at  Lachine  Rapids,  Canada 

(W.  L.  Walbank),  Western  Electrician  

23: 

24 

July  9,  1898 

Suggestions  Relative  to  Determining  Cost  of  Elec- 

tric Supply  (M.  E.  Turner),  Western  Electrician.  .  . 

23: 

143 

Sept.  10,  1898 

Cost  of  Power  (C.  S.  Brown),  Western  Electrician  .  „  . 

28: 

127 

Feb.  23,  1901 

Cost  of  Power  (C.  Grey),  Western  Electrician  

30: 

211 

March,  1902 

Cost  of  Power  (M.  J.  Eichorn),  Western  Electrician.  . 

31: 

69 

Aug.  2,  1902 

Graded  Costs  of  Electrical  Supply  (M.  E.  Turner), 

Western  Electrician  

35: 

204 

Sept.  10,  1904 

Some  Notes  on  the  Cost  of  Generating  Electrical 

Energy  (E.  J.  Fox),  Engineer  (London)  

27: 

219 

Feb.  26,  1904 

269 

March  11,  1904 

Effect  of  Load  on  the  Cost  of  Power  (E.  M.  Archi- 

bald), Engineer  (United  States)  

42: 

315 

May  1,  1905 

Cost  of  Power  in  Street  Railway  Service,  Machinery 

11: 

317 

Feb.,  1905 

726 


STEAM  POWER  PLANT  ENGINEERING 


ELECTRICAL  —  Continued. 
Cost  of  Electric  Power  at  the  Switchboard  (C.  H. 

Hile),  Power 25:  662  Nov.,  1905 

Power  Plant  Economies  (H.  G.  Stott),  Engineer 

(United  States) 43:  191  March  1,  1906 

Power  Costs  (Charles  E.  Lucke),  Electrical  Review 

(New  York) . . .. 50:  797  May  18,  1907 

Systems  of  Charging  for  Electricity  Supply  (W.  A. 

Toppin),  Electrical  Engineer  (London)  .........     39:  42  Jan.  11,  1907 

Rates  of  Charge  for  Electricity  and  Their  Effect 

on   Cost    (J.    S.    Codman),    Proceedings   of  the 

American  Institute  of  Electrical  Engineers 26:  31  April,  1907 

The  Principles  of  Modern  Rate-Making  for  Electric 

Light  and  Power,  Electrical  World 49:  1086          June  1,  1907 

Methods  of  Computing  Central  Station  Rates  in 

Boston,  Electrical  World 49:  1090          June  1,  1907 

The  Present  Tendency  of  Charging  for  Electricity 

(W.  A.  Toppin),  Electrical  Review  (London) ....     60:  945  June  7,  1907 

Electric    Power    Tariffs     (C.     S.     Nesey-Brown), 

Cassier's  Magazine 32:  304  Aug.,  1907 

The  Sale  of  Electricity  for  Lighting  Purposes  (L.  E. 

Bucknell),  Electrical  Engineer  (London) 40:  370  Sept.  13,  1907 

Rates  and  Systems  of  Charging  (Jacques),  Elec- 
trical Review  (London) 61:1074          Dec.  27,  1907 

GAS. 

Cost  of  Pumping  Station  in  New  York,  Electrical 

World 43:  820  April  30,  1904 

Comparative  Cost   of   Power  Generated  by  Steam 

Engine,  Water  Turbine,  and  Gas  Engine,  Engineer 

(London) 88:  320  Sept.  29,  1899 

Comparative  Cost  of  Generating  Power  by  Steam 

Engine,  Water  Turbine,  and  Gas  Engine  (J.  B.  C. 

Kershaw),  Engineering 70:  351  Sept.  14,  1900 

390  Sept.  21,  1900 

Cost  of  Gas  Power  for  Central  Station,  Engineer 71 :  27  Jan.  4,  1901 

Gas  Power  for  Central  Stations  (J.  R.  Bibbins), 

Engineering  Record 49:  11  Jan.  2,  1904 

Street  Railway  Journal 1089          Dec.  26,  1903 

Power 24:  100  Feb.,  1904 

Comparative    Cost    of    Steam    and    Gas    Plant, 

Engineering  Record 49:  310  March  5,  1904 

Is    Gas    Power    More    Economical    than    Water 

Power?  (H.  C.  T.  Horace),  Power 25:  599  Oct.,  1905 

Cost  of  Steam  Power  (Edit),  American  Electrician .  .     10:  114  March,  1898 

Cost  of  Steam  Electrical  Generating  Plant  (R.  C. 

Carpenter),  Electrical  World 43:  1016          May  28,  1904 

•  Economy    of    Power    Installations     (C.     Weiss), 

Engineering 66:  59  July  8,  1898 

Estimates  for  an  Electric  Light  Plant  in  New  York 

City,  Engineering  News 52:  583  Dec.  29,  1904 


FINANCE  AND  ECONOMICS  — COST  OF  POWER 


727 


GAS  —  Continued. 

Cost  of  a  Power  Station  in  Europe,  Street  Railway 

Journal 20:  210  Aug.  9,  1902 

Power   House   Cost,  Louisville   Electric    Railway, 

Street  Railway  Review 9:  592  Sept.,  1899 

MISCELLANEOUS. 

Improvements  in  Economy  of  the  Steam  Engine 

(W.  F.  Durand),  American  Electrician 11:  13  Jan.,  1899 

68  Feb.,  1899 

Cost  of  Power  (Edit),  Engineer  (London) 96:  285  Sept.  4,  1901 

Estimating  the  Cost  of  Power,  Engineer  (United 

States) 36:  285  Dec.  1,  1899 

Investigation  of  the  Cost  of  Power  (C.  G.  Gray), 

Engineer  (United  States) 39:  43  Jan.  1,  1902 

Efficient   Use   of   Steam   and   Labor  in   Isolated 

Plants  (P.  R.  Moses),  Engineering  Magazine 16:  99  Oct.,  1898 

Cost    Determination    in    Isolated    Plants    (P.    R. 

Moses),  Engineering  Magazine 20:  1082          March,  1901 

Cost  of  Pumping  at  a  Colliery   (R.   V.  Norris), 

Engineering  News 49:  228  March  12,  1903 

Economic   Power  Production    (R.   H.   Thurston), 

Engineering  Record 47:  35  Jan.  3,  1903 

Cost  of  Power  (Carpenter),  Power 24:  425  July,  1904 

Decreasing    Costs    in    the    Steam    Plant    (Edit), 

Engineer  (United  States) 42:  412  June  15,  1905 

The  Economy  of  a  Small-Sized  Coal  for  the  Power 

Plant  (P.  R.  Moses),  Engineering  Magazine 28:  783  Feb.,  1905 

Analysis  of  Central  Station  Costs  and  Revenues 

(H.  S.  Knowlton),  Engineering  Magazine 29:  238  May,  1905 

Cost  of  Operating  Buildings,  Engineering  Record 48:  759  Dec.  19,*  1903 

Power  Plant  Supervision  and  Accounting  (F.  W. 

Ballard),  Engineering  Record 51 :  687  June  17,  1905 

Power  Plant  Economy,  Power 25:  602  Oct.,  1905 

Relative   Efficiency   and   Desirability   of   Various 

Types  of  Engines  (A.  W.  Richter),  Street  Railway 

Review 10: 162  March  15,  1900 

STEAM. 

Comparative  Cost  of  Power  Generated  by  Steam 

Engine,     Water     Turbine,    and     Gas     Engine, 

Engineer  (London) 88:  322  Sept.  29,  1899 

Cost  of  Electric  Power  in  Comparison  with  Steam 

for  Traction,  Engineer  (London) 90:  600  Dec.  14,  1900 

Investigation    of    the    Cost    of    Power,    Engineer 

(United  States) 39:  161  March  1,  1902 

Cost    of    Steam    Power   per   Horse    Power   Year 

(J.  M.  Whitman),  Engineer  (United  States) 40:  741  Oct.  1,  1903 

Cost  of  Steam  Raising  ( J.  Holliday),  Engineering ...     68 :  739  Dec.  8,  1899 

808  Dec.  22,  1899 


728  STEAM  POWER  PLANT  ENGINEERING 

STEAM  —  Continued. 

Comparative  Cost  of  Generating  Power  by  Steam 

Engine,  Water  Turbine,  and  Gas  Engine  (J.  B.  C. 

Kershaw),  Engineering 70:  351  Sept.  14,  1900 

390  Sept.  21,  1900 

Cost  of  Steam  and  Electricity,  Engineering 74:  667  Nov.  21,  1902 

Comparative   Cost    of   Steam    and   Water   Power 

(W.  O.  Webber),  Engineering  Magazine 15:  923  Sept.,  1898 

Steam  Costs  in  an  Industrial  Combination  (W.  D. 

Ennis),  Engineering  Magazine 28:  86  Oct.,  1904 

Fuel   Economy   of  Engines  in  Electric   Railway 

Po wer  Stations  (Carpenter),  Engineering  News. . .     42:234  Oct.  12,  1899 

Reduction  in  Cost  of  Steam  Power  from  1870  to  1897, 

Engineering  Record » 37:  12  Dec.  4,  1897 

Economy  in  Use  of  Coal  for  Production  of  Power 

(I.  N.  Hollis),  Engineering  Record 46:  491  Nov.  22,  1902 

Economy    of    Fuel    in    Electric    Plants     (Edit), 

Engineering  Record 48:  233  Aug.  29,  1903 

Economical    Steam    Making    (Edit),    Engineering 

Record 48:  385  Oct.  3,  1903 

Cost  of  Fuel  and  Power  in  the  South,  Power 18: 13  Nov.,  1898 

Economical    Production    of    Steam    with    Special 

Reference  to  the  Use  of  Cheap  Fuel,  Power 19:  19  June,  1899 

Suggestions  for  Steam  Economy   (W.  M.   Kay), 

Engineer  (United  States) 42:  655  Oct.  2,  1905 

Yearly  Cost  of  One  Steam  Horse  Power,  Machinery .       9 :  374  March,  1903 

WATER. 
Cost  of  Water  and  Electric  Power  (G.  E.  Walsh), 

American  Electrician 16:  331  July,  1904 

Comparative  Cost  of  Power  Generated  by  Steam 

Engine,  Water  Turbine,  and  Gas  Engine,  Engineer 

(London) 88:  322  Sept.  29,  1898 

Investigation    of    the   Cost   of    Power,    Engineer 

(United  States) 39:  64  Jan.  15,  1902 

Comparative  Cost  of  Generating  Power  by  Steam 

Engine,  Water  Turbine,  and  Gas  Engine  (J.  B.  C. 

Kershaw),  Engineering 70:  351  Sept.  14,  1900 

390  Sept.  21,  1900 

Cost  of  Water  Power  in  France,  Engineering 76:  571  Oct.  23,  1903 

Comparative   Cost   of   Steam   and   Water   Power 

(W.  O.  Webber),  Engineering  Magazine 15:  923  Sept.,  1898 

Cost  of  Hydraulic  Transmission  of  Power  (E.  B. 

Ellington),  Engineering  Magazine 17:  233  May,  1899 

Cost  of  Hydraulic  Power  in  Switzerland,  Engineer- 
ing Record 41:  182  Feb.  24,  1900 

Analysis  of  the  Commerical  Value  of  Water  Power 

(A.  F.  Nagle),  Engineering  Record 46:540  Dec.  6,  1902 

Costs  of  Pumping  Water,  Power 20:  12  Nov.,  1900 


FINANCE  AND  ECONOMICS  —  COST  OF  POWER 


729 


COST  OF   POWER. 
(1908-1910.) 

Approximate  Cost  of  Gas  Power  (M.  P.  Cleghorn), 

Power  and  Engineer April  7,  1908 

Central  Station  vs.  Private  Plants,  Engineering Feb.  26,  1909 

Comparative  Cost  of  Power  Production,  Electrical  Age     40 :  63  March,  1909 

Electrical  World 53:  792  April  1,  1909 

Electrical  Review  and  Western  Electrician 55:  773  April  28,  1909 

Cost  of  a  Gas  Engine  and  of  a  Combined  Steam 

Plant,  Engineering  Record 60:  272  Sept.  4,  1909 

Cost  of  a  Kilowatt-Hour  (R.  A.  Day),  Electrical  World     54 :  853  Oct.  7,  1909 

Cost  of  Power  (H.  G.  Stott),  Pro.  Am.  Inst.  Elec. 

Engrs 28:  283  April,  1909 

Cost  of  Power  in  a  3000  Kw.  Turbine  Plant,  Elec- 
trical Review  and  Western  Electrician 55:  62  Oct.  2,  1909 

Cost  of  Power  in  a  1500  Kw.  Central  Station,  Engi- 
neering News 61 :  471  April  29,  1909 

Cost  of  Power  in  Small  Plants  (W.  E.  Snow),  Engi- 
neering Magazine 33:  169  May,  1908 

Cost  of  Power  in  Four  Central  Stations,  Electrical 

World 55:  813  March  31,  1910 

Cost  of  Power  for  Various  Industries  (C.  T.  Main), 

Jour.  Assoc.  Eng.  Soc 44:  151  March,  1910 

Engineering  Record 60:  711  Dec.  25,  1909 

Cost  of  Power  in  Varying  Units  (W.  O.  Webber), 

Engineering  Magazine 35:  562  July,  1908 

Cost  Systems  and  Time  Keeping,  Columbus,  O.,  Muni- 
cipal Electric  Lighting  Plant,  Engineering  News  Dec.  3,  1908 

Electric  Power  Costs  in  Small  Stations,  Engineering 

Record 59:  30  Jan.  9,  1909 

First  Cost  of  Plant  and  Cost  of  Generating  and 
Distributing  Electricity  for  Lights,  Brooklyn  Edi- 
son Co.,  Engineering  Contractor 33:  393  April  6,  1910 

Isolated  Power  Plant  Costs  and  Their  Relation  to 

Central  Station  Service(W.  F.  Lyod)  .ElectricalWorld     53 :  323  Feb.  4,  1909 

Isolated  Station  Records  and  Accounting  Power ....  April  28,  1908 

Operating  Costs  of  Large  Units,  Power  and  Engi- 
neer   31:  981  May  31,  1910 

Power  Costs  for  Factories,  Engineering  Record 60:  604  Nov.  27,  1909 

Power  Plant  Waste  (P.  R.  Moses),  Gassier 's  Maga- (  36:  497  Oct.,  1909 

zine 1  37:  12  Nov.,  1909 

Relation  of  Load  Factor  to  Power  Costs,  Jour.  Wes. 

Soc.  Engr 14:  241  April,  1909 

Engineering  Record 59:  702  June  5,  1909 

Representative   Data   from   Electric   Power  Plant 

Operation  (H.  S.  Knowlton),  Engineering  Magazine     36 :  833  Feb.,  1909 

Systems  of  Charging  for  Electrical  Energy  (W.  T. 

Ryan),  Engineering  Magazine April,  1909 

The  Small  Station  and  its  Economical  Operation, 

Western  Electrician 43:  10  July  4,  1908 

The  Valuation  of  Steam  Power  Plants  (C.  T.  Main), 

Electrical  Age 39 :  228  Oct.,  1908 

Useful  Figures  from  Practical  Power  Plant  Opera- 
tion, Electrical  World 54:  781  Sept.  30,  1909 

Working  Results  from  a  Gas-Electric  Power  Plant 

(J.  R.  Bibbins),  Pro.  Am.  Inst.  Elec.  Engrs 27:  1123          July  1,  1909 


CHAPTER  XVIII. 

TESTING  AND  MEASURING  APPARATUS. 

403.  General.  —  The  importance  of  maintaining  a  system  of  records 
has  been  referred  to  in  paragraph  388.     The  various  items  which  may 
be  recorded  and  the  instruments  and  appliances  used   in   this   con- 
nection are  outlined  in  the  accompanying  chart.     In  large  stations  a 
full  complement  of  indicating,  recording,  and  integrating  instruments 
may  prove  to  be  a  good  investment  if  intelligently  and  closely  studied 
by  the  operating  engineer  with  a  view  to  locating  and  eliminating 
unnecessary  losses.     The  instruments  should  be  inspected  and  cali- 
brated at   intervals,  since    many  of  them  are  delicately  constructed 
and  are  apt  to  become  inaccurate  after  a  few  months'  service.     Steam 
gauges,  thermometers,  and  pyrometers,  and  particularly  water  meters 
are  subject  to  appreciable  error  after  considerable  use.     Voltmeters, 
ammeters,  and  other  switchboard  instruments  are  easily  deranged,  espe- 
cially when  subjected  to  continuous  vibration  or  to  high  temperature. 

404.  Weighing    the    Fuel.  —  In    most    small    plants    the    delivery 
tickets  of  the  coal  dealer  are  depended  upon  for  the  weight  of  coal 
used,  no  attempt  being  made  to  determine  its  evaporative  value,  and 
the  economy  of  the  plant  is  judged  by  the  size  of  the  coal  bill.     In 
such  cases  a  considerable  saving  can  be  effected  by  keeping  a  daily 
record  covering  at  least  the  coal  and  water  consumption.     The  coal 
can  be  conveniently  weighed  on  ordinary  platform  scales.     In  a  num- 
ber of  large  stations  the  weight  of  coal  is  determined  by  suspended 
weighing  hoppers,  which  may  be  stationary,  as  in  Fig.  109,  or  mounted 
on  a  traveling  truck,  as  in  Fig.  110.     The  scales  of  such  devices  are 
made  indicating,   autographic,  integrating,  or   a   combination   of  the 
three,  the  latter  costing  but  little  more  than  the  simple  indicating  or 
recording  devices. 

405.  Measurement  of  Water.  —  The  most  accurate  means  of  measur- 
ing water  is  by  the  use  of  two  or  more  tanks  resting  upon  scales, 
arranged  to  be  filled  and  emptied  alternately.     This  method,  however, 
involves  considerably  more   time  than  is   ordinarily  at  the  disposal 
of  the  fireman  or  engineer.     The  usual  practice  is  to  place  a  hot-water 
meter  on  the  pressure  side  of   the  feed  pump,   with  provision  for 

calibration  without  shutting  off  the  feed  supply  to  the  boilers. 

730 


TESTING  AND  MEASURING  APPARATUS 


731 


Weights, 


Pressures. 


Tempera- 
tures 


Power . 


Flue  gas 
analysis 


Moisture. 


Fuel  analysis 


TESTING  AND  MEASURING  APPARATUS 
STEAM   PLANT. 


Fuel 


Water 


•Platform  scales,  indicating  and  autographic. 
Suspension   hoppers,   indicating  and   auto- 

graphic. 
Platform  scales  and  tanks. 

Piston....  ^ 

Rotary  .  .  .  ^Integrating. 


Water  meters   - 


Disk  .....  J 


Steam 


Venturi,     indicating 

autographic. 
Weirs. 
Weighing  condensed  steam. 

St*  Jolm's>  autographic. 
Burnham^  indicating. 


and 


(  Bourdon  gauge,  indicating  and  autographic. 
ga  ............   (Manometers,  mercurial,  indicating. 


/Manometers  —  mercurial,     indicating,     and 

autographic. 

Low -[  Manometers  —  water,  indicating,  and  auto- 
graphic. 

Diaphragms,  indicating  and  autographic, 
f  Mercurial  thermometers,  indicating. 
Up  to  800  deg.  F.  1  Expansion    thermometers,    indicating    and 

autographic. 
Expansion    thermometers,    indicating    and 

autographic. 

onn  *    otnn  j        v      Resistance    thermometers,    indicating    and 
800  to  2500  deg.  F.         autOgraphic. 

Thermo-electric    thermometers,     indicating 

and  autographic. 

Optical    pyrometer,    indicating    and    auto- 
Over  2500  deg.  F.        graphic. 

Platinum  or  clay  ball  pyrometer. 

Indicators,  hand  manipulated. 

Indicators,  continuous  autographic. 
'Rope  brake. 

Prony  brake. 

Absorption  dynamometers. 

.Electric  generator. 
fOrsat  apparatus. 
I  Arndt's  econometer,  indicating. 
I  Sarco  and  Ados  recorder,  autographic. 
LUehling  gas  composimeter,  autographic. 
Tin  air Hygrometer,  indicating  and  autographic. 

|In  steam Calorimeters. . . .  j|g^^ 

rMahler  bomb. 
Coal  calorimeters.     «- 

LParr. 
^Gas  calorimeter Junker. 

ELECTRICAL   PLANT. 


Indicated 


Developed. 


{ 


Voltage Voltmeters,  A.C.  and  D.C.,  indicating  and  autographic. 

Current Ammeters,  A.C.  and  D.C.,  indicating  and  autographic. 

Output Wattmeters,  A.C.  and  D.C.,  integrating  and  autographic. 

Power  factor  ..Power  factor  meters,  A.C.  only,  indicating  and  autographic. 

Frequency Frequency  meter,  A.C.  only,  indicating. 

Synchronism  .  .Synchronizers,  A.C.  only,  indicating. 


732 


STEAM  POWER  PLANT  ENGINEERING 


There  are  several  types  of  meters  in  common  use.  Fig.  419  illus- 
trates the  piston  type,  in  which  reciprocating  pistons  are  displaced  by  a 
definite  volume  of  water;  Fig.  305,  the  rotary  type,  depending  upon 


FIG.  419.    A  Typical  Piston  Water  Meter. 

the  displacement  of  rotating  impellers;  Fig.  420,  the  disk  type,  in 
which  impellers  are  given  a  combined  rotating  and  tilting  motion. 
When  periodically  calibrated,  water  meters  give  satisfactory  results. 

When  graduated  to  read 
in  pounds  the  accuracy  de- 
pends upon  the  temper- 
ature range  of  the  water; 
thus  the  density  of  water 
a£  62  degrees  F.  is  62.36 
pounds  per  cubic  foot  and 
at  212  degrees  it  is  59.76, 
a  range  of  2.6  pounds  per 
cubic  foot.  Hence  a  meter 
calibrated  to  read  correctly 
at  62  degrees  F.  will  have 
an  error  of  about  4.2  per 
cent  if  used  to  measure 
water  at  212  degrees.  The 
average  range  in  feed  tern- 


F».  420.    A  Typical  Disk  Water  Meter. 

than  40  degrees,  and  if  the  meter  is  calibrated  for  the  mean  tem- 
perature the  error  is  somewhat  less  than  one  per  cent. 

The  Venturi  meter,  Fig.  421,  is  frequently  employed  for  measuring 
large  volumes  of  water,  as  in  city  waterworks,  and  in  connection  with 
condensing  plants.*  It  amounts  practically  to  a  constriction  in  the 
diameter  of  the  pipe,  is  readily  installed,  and  the  total  absence  of 

*  Tests  on  a  Venturi  Meter  for  Boiler  Feed.  Proc.A.S.M.E.,Mid.-Oct.,  1909,p.l065. 


TESTING  AND  MEASURING  APPARATUS 


733 


working  parts  is  a  great  advantage.  The  meter  is  supplied  with 
either  indicating  or  autographic  manometer.  With  water  at  constant 
temperature  the  error  in  the  readings  should  not  exceed  one  per 
cent. 

The  pitometer  is  a  simple  adaptation  of  the  well-known  pitot  tube, 
and  is  used  for  measuring  the  flow  of  water  through  pipes  where  it  is 


BrJ 


FIG.  421.     Principles  of  the  Venturi  Meter. 


MANOMETER 


impracticable  to  insert  a  meter.  It  is  only  necessary  to  drill  a  small 
hole  in  the  pipe  for  the  introduction  of  the  tube.  The  volume  flowing 
may  be  calculated  from  the  readings  of  the  manometer  or  may  be 
autographically  recorded. 

In  measuring  large  volumes  of  water  flowing  in  open  channels  the 
measurements  are  made  by  weirs  of  suitable  proportions  or  by  current 
meters. 

Water  Measurement,  General:   Eng.  Rec.,  Feb.  15,  1902. 

Water  Meters:  Trans.  A.S.M.E.,  18-134,  14-676,  5-63;  Engng.  News,  Jan.  3, 
1907,  March  9,  1905,  June  16,  1904,  p.  569;  Eng.  Rec.,  Nov.,  1903;  Stevens  Ind., 
Jan.,  1901;  Jour.  New  Eng.  Waterworks  Assn.,  June,  1907;  Elecn.,  Lond.  May 
8,  1908. 

Venturi  Meter:  Revue  Technique,  Feb.  10,  1905,  Eng.  News,  Feb.  28,  1901; 
Prac.  Engr.,  Feb.  15,  1907;  Pro.  A.S.M.E.,  Nov.,  1906;  Trans.  A.S.C.E.,  Nov., 
1907,  57-531  (1906);  Engng.,  Feb.  22,  1907,  p.  236. 

Pitometer:  Jour.  Franklin  Inst.,  Dec.,  1907,  p.  425;  Tech.  Quar.,  June,  1907; 
Jour.  New  Eng.  Waterworks  Assn.,  June,  1906;  Trans.  A.S.M.E.,  25-184;  Sib.  Jour. 
Engng.,  May,  1902;  Jour.  Assn.  Eng.  Soc.,  Aug.,  1901;  Eng.  News,  March  31,  1904, 
Dec.  21,  1905;  Pro.  Engrs.  Soc.  of  West.  Penn.,  Dec.,  1906. 

Pitot  Tube:  Am.  Mach.,  Aug.  9,  1906,  p.  175;  Trans.  A.S.C.E.,  47-6,  57-265, 
25-184;  Eng.  News,  March  31,  1904,  p.  318,  Dec.  21,  1905,  p.  660;  Progressive  Age, 
June,  1906,  p.  63;  Jour.  Assn.  Eng.  Soc.,  Aug.,  1901,  p.  35;  Sib.  Jour.,  May,  1902; 
Cal.  Jour,  of  Tech.,  May,  1905. 

Weirs:  Engr.  (Lond.),  June  5,  1903,  p.  562,  Aug.  17,  1906;  Am.  Soc.  Civ.  Engrs., 
44-160;  Eng.  Rec.,  July  13,  1901,  p.  32. 


734 


STEAM  POWER  PLANT  ENGINEERING 


Current  Meters:  Cal.  Jour,  of  Tech.,  April,  1904;  Eng.  News,  March  7,  1907, 
p.  263,  Feb.  12,  1902;  Eng.  Rec.,  Dec.  19,  1903;  Pro.  Am.  Soc.  Civ.  Engrs.,  Sept., 
1901,  Nov.,  1901,  Dec.,  1901. 

Piezometers:  Power,  Aug.,  1907,  p.  569;  Pro.  Am.  Soc.  Civ.  Engrs.,  44-34,  49-112, 
51-252;  Eng.  News,  Sept.  13,  1906,  p.  271;  Power,  Aug.,  1907,  p.  569. 

406.  Measurement  of  Steam.  —  In  surface-condensing  plants  the 
weight  of  steam  consumed  by  the  engines  is  conveniently  obtained 
by  weighing  the  condensed  steam;  in  practice,  however,  this  method 


Fio.  422.    St.  John's  Steam  Meter. 


FIG.  422a.    Principles  of  the  Burnham 
Steam  Meter. 


is  adopted  only  when  testing  the  plant,  the  feed-water  measurement 
sufficing  for  general  recording  purposes. 

When  steam  is  supplied  to  various  points  and  the  weights  cannot  be 
readily  determined  by  condensing,  steam  meters  are  sometimes  used. 
The  St.  John's,  Fig.  422,  is  the  best  known  in  this  country.  This 
apparatus  records  the  weight  of  steam  passing  through  the  seat  of  an 


TESTING  AND  MEASURING  APPARATUS 


735 


automatically  lifting  valve  which  rises  and  falls  as  the  demand  for 
steam  increases  or  diminishes.  When  the  maximum  fluctuation  in 
steam  pressure  is  less  than  10  pounds  per  square  inch  and  the  moisture 
in  the  steam  is  practically  constant,  this  apparatus  is  said  to  register 
within  two  per  cent  of  the  actual  weight  flowing. 

The  Burnham  steam  meter,  Fig.  422a,  a  recently  patented  device, 
offers  the  advantages  of  low  first  cost  and  simplicity  of  installation. 
This  apparatus  is  attached  to  the  steam  pipe  in  a  manner  similar  to  a 
simple  hydrostatic  lubricator  and  occupies  about  the  same  space. 
It  is  based  on  the  principle  of  the  pitot  tube,  and  the  weight  of  steam 
flowing  per  unit  of  time  is  read  from  a  graduated  scale.  Half-inch 
pipe  fittings  are  used  in  connecting  up. 

In  Europe  the  principles  of  the  Venturi  meter  have  been  success- 
fully applied  to  the  measurement  of  steam.  jfGluckauf,  Dec.  9,  1905.) 

Steam  Meters:  Proc.  A.S.M.E.,  Mid.-Nov.,  1909,  p.  1239. 


DRAFT  GAUGE 


FIG.  423.     Different  Forms  of  Manometer  Pressure  Gauges. 

406a.  Pressure  Gauges. — The  Bourdon 
type  of  gauge,  either  autographic  or  indi- 
cating (Fig.  424),  is  the  most  familiar  and 
satisfactory  means  of  measuring  pressures 
up  to  1500  pounds  per  square  inch  or 
more,  although  diaphragm  gauges  are  also 
used  and  both  are  employed  as  vacuum 
gauges.  For  the  latter  purpose,  however, 
the  mercurial  vacuum  gauge  has  the  ad- 
vantage of  greater  accuracy  and  is  not 
subject  to  derangement.  Bourdon  gauges 
should  be  frequently  standardized  by  com- 
parison with  a  gauge  of  known  accuracy,  a  mercury  column,  or  a  gauge 
tester. 

For  measuring  very  low  pressures,  such  as  are  found  in  boiler  flues 


FIQ.  424.    Bourdon  Pressure 
Gauge. 


736  STEAM   POWER  PLANT  ENGINEERING 

or  gas  mains,  indicating  or  recording  diaphragm  gauges  may  be  had,  but 
some  form  of  U  tube  manometer  is  generally  employed,  the  design  best 
adapted  to  the  purpose  depending  upon  the  accuracy  required.  The 
simple  U  tube  (Fig.  423)  when  filled  with  mercury  may  be  used  for 
pressures  limited  only  by  the  inconvenience  due  to  length  of  tubes,  or; 
with  water  as  the  fluid,  for  pressures  only  a  fraction  of  an  ounce  per 
square  inch.  Where  greater  accuracy  is  required  than  can  be 
obtained  with  the  simple  U  tube,  some  modification  may  be  employed, 
such  as  the  Eames  draft  gauge  with  one  inclined  leg  which  magnifies 
the  reading  several  times.  A  form  of  sensitive  gauge  is  sometimes 
used  which  depends  upon  the  use  of  two  fluids  of  different  specific 
gravity,  as  oil  and  water. 

Pressure  Gauges,  General  References:  Mech.  Engr.,  Aug.  17,  1907;  Am.  Elecn., 
July,  1901;  Engng.,  Aug.  23,  1907;  Elec.  World,  Feb.  2,  1907,  p.  258;  Power,  March, 
1905,  p.  184. 

Recording  Pressure  Gauge:  Trans.  A.S.M.E.,  11-225,  14-325;  Elec.  World, 
April  28,  1906,  p.  886. 

Draft  Gauges:  Trans.  A.S.M.E.,  21-123;  Engr.  U.S.,  Feb.  15,  1907,  p.  218; 
Mech.  Engr.,  Oct.  27,  1906. 

407.  Measurement  of  Temperature.  —  For  power-plant  purposes 
mercurial  thermometers  are  most  convenient  for  measuring  tempera- 
tures up  to  400  degrees  F.,  and  are  inexpensive.  For  higher  tempera- 
ture, up  to  say  800  degrees  F.,  they  are  also  adapted,  but  must  be 
made  of  special  glass  and  the  space  above  the  mercury  filled  with 
nitrogen  under  pressure  to  prevent  vaporization  of  the  mercury.  Such 
thermometers  must  be  used  intelligently,  and  should  be  standardized 
from  time  to  time,  since  they  are  subject  to  considerable  change. 
The  Bureau  of  Standards  at  Washington,  B.C.,  is  prepared  to  furnish 
certificates  for  which  a  nominal  charge  is  made. 

Fig.  425  shows  a  form  of  thermometer  which  is  much  used  where  a 
continuous  autographic  record  is  required.  It  depends  for  its  oper- 
ation upon  the  pressure  produced  by  a  fluid,  liquid  or  gaseous, 
contained  in  a  small  bulb  and  exposed  to  the  temperature  to  be 
measured.  The  pressure  is  transmitted  to  the  recording  mechanism 
through  a  flexible  capillary  tube  which  may  be  of  considerable  length. 
Such  thermometers  are  suitable  for  feed  water,  flue  gas,  and  tempera- 
tures not  exceeding  1000  degrees  F. 

Fig.  426  illustrates  a  form  of  electrical  pyrometer  employing  thermo- 
couples which  has  come  into  wide  use  as  a  reliable  means  of  measur- 
ing temperatures  up  to  2600  degrees  F.  The  couples  most  frequently 
used  are  composed  of  platinum  and  platinum-rhodium,  platinum  and 
platinum-iridium,  copper  and  copper-constantan,  and  copper  and  nickel, 


TESTING  AND  MEASURING  APPARATUS 


737 


the  first  named  being  adapted  to  the  higher  ranges  of  temperature.  The 
electro-motive  force  set  up,  when  the  thermo-j  unction  is  heated,  is  pro- 
portional to  the  temperature  and  is  measured  by  means  of  a  sensi- 


FIG.  425.    Bristol  Recording  Pyrometer. 


FIG.  426.    Bristol  Thenno-EIectric  Pyrometer. 

tive  millivoltmeter  which  is  usually  graduated  to  read  temperature 
directly.  Thermo-couples  may  be  made  to  give  an  autographic  record 
by  means  of  a  thread  recorder. 


738 


STEAM  POWER  PLANT  ENGINEERING 


Fig.  427  shows  the  element  of  an  electrical  thermometer  based  upon 
the  change  in  resistance  of  a  platinum  wire  when  subjected  to  change 
in  temperature.  The  resistance,  in  terms  of  temperature,  is  measured 
by  a  Whipple  indicator,  a  convenient  and  portable  form  of  Wheat- 


PLATINUM  COII. 


4        PLATINUM        WIRE       LEADS 


FIG.  427.     Element  for  Callendar  Resistance  Pyrometer. 


stone  bridge,  or  may  be  autographically  recorded  by  means  of  a  Callen- 
dar recorder.  Resistance  thermometers  of  this  type  are  very  sensitive 
and  accurate,  not  easily  deranged,  and  are  limited  in  range  only  by  the 
fusing  points  of  the  platinum  and  the  porcelain  protecting  sheath. 

For  higher  temperatures  and  for  obtaining  the  temperatures  of 
inclosed  spaces  above  about  900  degrees  F.,  such  as  boiler  furnaces, 
annealing  ovens,  and  kilns,  various  forms  of  optical  and  radiation 
pyrometers  have  been  devised.  In  such  devices  no  part  of  the  instru- 


DirFUSING  GLASS 


FLAME 
GAUQE 


AMYL-ACETAT 
LAMP 


FIG.  428.    Wanner  Optical  Pyrometer  in  Position  for  Standardizing. 

ment  is  exposed  to  the  temperature  to  be  measured  and  hence  suffers 
no  injury  from  this  cause.  Optical  pyrometers  are  based  upon  the 
measurement  of  the  brightness  of  the  hot  body  by  comparison  with 
a  standard.  The  Wanner  optical  pyrometer  is  shown  in  Fig.  428. 


TESTING  AND  MEASURING  APPARATUS 


739 


After  standardizing  by  comparison  with  an  amyl-acetate  lamp,  it  is 
only  necessary  to  focus  the  instrument  upon  the  source  of  heat  to  be 
measured  and  the  temperature  is  read  on  the  graduated  scale. 


TABLE  109. 


TYPES  OF  THERMOMETERS  IN  GENERAL  USE. 


Principle  of  Operation. 

Type. 

Range  in  Degrees  F. 
for  which  they 
can  be  used. 

Gas  

—  400  to  +2900 
-35  to  +950 

—  325  to  +100 
0  to         950 

0  to     2900 
—  400  to  +2900 

—  400  to  +2200 

300  to  4000 
—  400  to  Sun 

11  00  to  Sun 
32  to  3000 

32  to    3350 

Transpiration  and 
cosity. 

Thermo-electric 

change  in  volume  or 
length  of  a  body  with 
temperature. 

vis-  Those  depending  on  the 
flow  of  gases  through 
capillary  tubes  or  small 
apertures. 
.  .      Those  depending  on  the 

Mercury,    Jena     glass, 
and  nitrogen 
Glass  and  petrol  ether. 

Unequal  expansion  of 
metal  rods. 
The  Uehling  

Galvanometric  . 

Electric  resistance. 
Radiation  

electro-  motive       force 
developed  by  the  dif- 
ference in  temperature 
of  two  similar  thermo- 
electric  junctions    op- 
posed to  one  another. 
....Those  utilizing  the  in- 
crease in  electric  resist- 
ance   of  a   wire    with 
temperature. 

Direct  reading  on  indi- 
cator   or  bridge  and 
galvanometer. 

Thermo-couple  in  focus 
of  mirror. 
Bolometer  

Optical  

heat  radiated  by  hot 
bodies. 

Photometric   compari- 
son. 
Incandescent    filament 
in  telescope. 
Nicol  with  quartz  plate 
and  analyzer. 
Platinum      ball    with 
water  vessel. 

Alloys  of  various  fusi- 
bilities. 

Calorimetric  

change  in  the  bright- 
ness or  in  .the  wave 
length  of  the  light 
emitted  by  an  incan- 
descent body. 

Fusion  

specific  heat  of  a  body 
raised  to  a  high  tem- 
perature. 

unequal  fusibility  of 
various  metals  or 
earthenware  blocks  of 
varied  composition. 

Radiation  pyrometers  depend  upon  the  measurement  of  the  heat 
radiated  from  the  hot  body.    The  Fery  radiation  pyrometer,  Fig.  429, 


740 


STEAM  POWER  PLANT  ENGINEERING 


is  the  best-known  instrument  of  this  type.  When  focused  upon  the 
source  of  heat  a  cone  of  rays  of  definite  angle  is  reflected  by  means  of 
the  mirror  upon  a  thermo-couple  located  in  its  focus.  The  electro- 
motive force  set  up  is  measured  in  terms  of  the  temperature  of  the 


TO  GALVANOMETER 


FIG.  429.    Fe"ry  Radiation  Pyrometer. 

source  of  heat  by  a  millivoltmeter.  Neither  the  couple  nor  any 
part  of  the  instrument  is  ever  subjected  to  a  temperature  much  above 
150  degrees  F.  The  indications  are  practically  independent  of  the 
distance  from  the  source  of  heat,  and  the  range  is  without  limit. 

Table  109  embodies  in  outline  the  principles  and  temperature  ranges 
of  the  various  types  of  thermometers  in  use.  Temperature  ranges 
verified  by  U.  S.  Bureau  of  Standards. 

Indicating  and  Recording  Thermometers,  Expansion  Type:  Sci.  Am.  Sup., 
Dec.  16,  1905;  Trans.  A.S.M.E.,  22-143;  Jour.  Am.  Chem.  Soc.,  16-396;  Jour.  Soc. 
Chem.  Ind.,  13-61 ;  Philosoph.  Mag.,  50-251,  1900. 

Indicating  and  Recording  Pyrometers,  Thermo-Electric:  Jour.  West.  Soc.  Engrs., 
Sept.,  1907;  Cassier's  Mag.,  Aug.,  1905;  Elec.  Rec.,  Jan.  12,  1901;  Elec.  Chem.  and 
Met.,  June,  1901. 

Indicating  and  Recording  Pyrometers,  Electric-Resistance:  Engng.,  May,  1899; 
Jour.  Chem.  Soc.,  1890,  1895;  Jour.  Iron  and  Steel  Inst.,  1892;  Pro.  Royal  Inst., 
Vol.  XVI,  1901;  Bureau  of  Standards,  3-641,  1907;  Electrician,  March  17,  1905, 
p.  880. 

Indicating  and  Recording  Pyrometers,  Optical:  Elecn.,  Lond.,  Aug.  17,  1906; 
Am.  Mach.,  Vol.  28,  160-29;  Sch.  of  Mines  Quarterly,  April,  1907;  Bulletin  No.  2, 
Bureau  of  Standards,  Wash.,  1905;  Jour,  de  Phys.,  Sept.,  1904;  Engng.,  Sept.  6,  1907, 
Oct.  18,  1907;  Cal.  Jour,  of  Tech.,  Aug.,  1907;  Bureau  of  Standards.  Bulletin  No.  2; 
Iron  Age,  73-24. 

Miscellaneous  References:  Engng.  Times,  March,  1904;  Engng.,  Feb.  17,  1903, 
March  6,  1904;  Sci.  Am.  Sup.,  July  22,  1905;  Min.  Rept.,  Aug.  8,  1901;  Iron  Age, 
Feb.  7, 1907 ;  Iron  and  Coal  Tds.  Rev.,  May  10,  1907 ;  Am.  Elecn.,  May,  1904 ;  Physical 
Rev.,  8-193;  Roy.  Soc.  of  Lond.,  66-86,  1900. 


TESTING  AND  MEASURING  APPARATUS 


741 


408.  Power  Measurements.  —  The  indicated  horse  power  of  recip- 
rocating engines  is  usually  obtained  by  means  of  the  steam-engine 
indicator.  There  are  several  reliable  types  to  be  had,  including  the 
continuous  indicator,  which  permits  of  several  diagrams  being  taken 
successively  on  the  same  paper.  Among  other  devices  may  be 
mentioned  mean  pressure  indicators  and  those  giving  the  horse  power 
directly. 

The  developed  horse  power  is  determined  by  some  form  of  absorp- 
tion dynamometer.  For  description  of  such  dynamometers  see  Appen- 
dix C,  article  XV,  A.S.M.E.  code  for  conducting  steam-engine  tests. 

Power  Measurements:  Trans.  A.S.M.E.,  13-531;  Am.  Mach.,  Vol.  30,  No.  27, 
Vol.  31,  No.  5;  Mechanical  Engr.,  Feb.  23,  1907;  Engng.,  June  14,  1907,  p.  768. 

Indicators,  Continuous:   Trans.  A.S.M.E.,  18-1020;  Power,  Jan.,  1907,  p.  26. 

Prony  Brakes:  Trans.  A.S.M.E.,  15-62;  Am.  Mach.,  July  27,  1905,  p.  127;  Eng. 
News,  Vol.  44,  p.  216. 

Water  Absorption  Dynamometers:  Prac.  Engr.,  Sept.  14,  1906;  Trans.  A.S.M.E., 
11-958;  Eng.  News,  Vol.  51,  p.  475;  Prac.  Engr.,  Sept.  14,  1906,  p.  326. 


FIG.  430.    Orsat  Apparatus  for  Flue  Gas  Analysis. 

409.   Flue-Gas  Analysis.  —  The  simplest  device  for  the  analysis  of 
flue  gases   is  the  Orsat  apparatus   (Fig.  430).     In   this   apparatus  a 


742' 


STEAM   POWER   PLANT   ENGINEERING 


measured  volume,  representing  an  average  sample  of  the  gas,  is  forced 
successively  through  pipettes  containing  solutions  of  caustic  potash, 
pyrogallic  acid,  and  cuprous  chloride  in  hydrochloric  acid,  respectively, 
thus  removing  the  carbon  dioxide,  the  oxygen,  and  the  carbon  monox- 
ide, the  contraction  of  volume  being  measured  in  each  case. 

Orsat  Apparatus:  Trans.  A.S.M.E.,  18-901 ;  Steam  Boilers,  Peabody  and  Miller, 
Chap.  II;  Power,  Aug.,  1907,  p.  532;  Engr.  U.S.,  Jan.  1,  1907,  p.  71. 

For  most  practical  purposes  it  is  sufficient  to  determine  the  carbon 
dioxide.  A  simple  and  efficient  device  for  continuously  indicating 
the  per  cent  of  CO2  is  Arndfs  econometer,  Fig.  431.  This  apparatus 


FIG.  431.     Arndt's  Econometer. 

is  a  gas-weighing  balance  consisting  essentially  of  a  sensitive  beam 
from  one  end  of  which  is  suspended  a  glass  globe,  closed  at  the  top  and 
open  at  the  bottom,  of  about  one  pint  capacity,  and  from  the  other 
end  a  compensating  rod  and  scale  pan.  When  not  in  operation  the 
globe  is  filled  with  air  and  the  scale  pan  and  globe  are  in  perfect 
balance,  the  indicator  pointing  to  zero.  When  in  operation  the  flue 
gases,  thoroughly  dried  and  filtered,  are  introduced  in  a  continuous 
flow  into  the  body  of  the  hollow  globe  by  means  of  a  glass  tube. 
The  larger  the  per  cent  of  CO2  present  in  the  contents  of  the  globe 
the  greater  will  be  the  deflection  of  the  pointer,  since  CO2  is  about 


TESTING  AND  MEASURING  APPARATUS 


743 


50  per  cent  heavier  than  atmospheric  air.  The  scale  is  graduated 
to  read  from  0  to  21  per  cent  CO2,  and  the  results  obtained  check 
closely  with  those  of  the  Orsat  apparatus.  Arndt's  econometer  is 
not  portable,  though  it  may  be  placed  almost  anywhere  where  it  can 
be  easily  seen  by  the  fireman.  When  there  are  a  number  of  boilers, 
and  it  is  not  desired  to  have  a  separate  instrument  for  each,  the  eco- 
nometer is  connected  with  the  breeching  of  each  boiler  by  suitable  piping, 
the  gas  from  one  boiler  at  a  time  being  analyzed. 

For  descriptive  details  see  circular  issued  by  Joseph  Wickes,   106 
Fulton  Street,  New  York. 


The  Ados  recorder,  Fig.  432,  Sarco  recorder,  Fig.  433,  and  Uehling 
composimeter  are  well-known  types  of  instruments  which  give  con- 
tinuous autographic  records  of  the  percentage  of  CO2  on  clock-driven 
charts.  These  devices  are  very  satisfactory,  but  are  rather  expensive 
and  usually  beyond  the  appropriation  of  small  boiler  plants. 

The  Sarco  C02  recorder  operates  as  follows:  Gas  is  drawn  con- 
tinuously from  the  flue  through  a  f-inch  pipe  by  rneans  of  the  aspi- 


744 


STEAM  POWER  PLANT  ENGINEERING 


rator  Q,  and  enters  the  apparatus  through  the  tube  D,  flowing  as  indi- 
cated by  the  arrows  through  C  and  E.  The  aspirator  requires  from  6 
to  8  gallons  of  water  per  hour,  which  is  discharged  into  the  pressure- 
regulating  tank  L.  The  greater 
portion  overflows  through  R, 
and  the  balance  is  caused  to 
issue  in  a  fine  stream  by  adjust- 
ment of  cock  S  into  H  and  G 
and  thence  to  an  upper  com- 
partment of  vessel  K,  which  it 
gradually  fills,  compressing  the 
air  which  it  contains  and  trans- 
mitting this  pressure  to  the 
lower  chamber  through  a  con- 
necting tube.  The  lower  cham- 
ber is  filled  with  a  solution  of 
1  part  glycerine  to  3  parts 
water,  which  is  driven  out  into 
the  calibrated  vessel  C.  When 
it  has  risen  to  the  height  of  the 
inlet  and  outlet  the  flow  of  gas 
is  interrupted  and  a  portion  is 
trapped.  By  the  time  the  lower 
end  of  the  center  tube  in  C  is 
sealed  off  the  excess  of  gas  has 
been  forced  out  against  the 
slight  resistance  of  the  elastic 
bag  P,  and  the  vessel  contains 
exactly  100  c.c.  at  atmospheric 
pressure.  During  this  time  the 
aspirator  is  drawing  the  flue  gas 
through  the  seal  F.  As  the 
liquid  rises  higher  in  (7,  the  en- 
trapped gas  is  forced  through 
the  small  tube  Z  into  A,  which 
is  filled  with  a  solution  of  caus- 
tic potash.  Here  the  CO2  is  absorbed  and  the  potash  solution 
is  forced  into  vessel  B,  which  has  an  outer  jacket  filled  with 
glycerine  supporting  a  float  N  suspended  from  the  pen- gear  M. 
A  thin  tube  through  the  float  keeps  the  air  in  B  at  atmos- 
pheric pressure.  As  the  float  rises  the  pen-lever  swings  upward, 
carrying  the  pen  Y,  which  makes  a  vertical  line  upon  the  clock-driven 


FIG.  433.    Sarco  CO2  Recorder. 


TESTING  AND  MEASURING  APPARATUS 


745 


recording  drum.  The  quantity  of  potash  solution  displaced,  and  con- 
sequently the  lift  of  the  float,  is  dependent  upon  the  amount  of  CO2 
absorbed  from  the  100  c.c.  of  flue  gas.  The  chart  is  graduated  to 
indicate  directly  the  percentage  absorbed.  By  the  time  this  opera- 
tion has  been  completed  water  has  rilled  tank  K  and  risen  into  the 
siphon  G,  which,  upon  starting,  very  rapidly  empties  the  tank  and 
allows  the  liquid  from  C  to  return  to  the  lower  compartment.  The 
float  returns  to  its  original  position  and  the  remaining  gas  passes  out 
through  E. 

Ados  CO2  Recorder:   Engng.,  Jan.  12,  1906;  Sci.  Am.  Sup.,  Dec.  22,  1906. 
Sarco  CO2  Recorder:   U.S.  Engr.,  Nov.  1,  1907,  p.  1001. 
Uehling  Composimeter:    Power,  June,  1907,  p.  404. 
American  CO2  Indicator:   Power,  Dec.,  1907. 

Flue  Gas  Analysis,  Miscellaneous  Apparatus:   Power,  April,  1907,  p.  243;  Engr. 
U.S.,  Jan.  1,  1907,  p.  71;  Elecn.,  Lond.,  Nov.  16,  1906. 

410.   Moisture  in  Steam.  —  Several  forms  of  calorimeters  are  avail- 
able for  determining  the  quality  of  steam.     The  simplest   as  well   as 
the  most  satisfactory,  if  the  percentage  of  entrained  moisture  is  not 
beyond    its    range,    is    the    throttling 
calorimeter,  Fig.  435.      In  this  device 
the  sample  of  steam,  which  is  taken 
from  the  steam  pipe  by  means  of  the 
perforated  nipple,  is  allowed  to  expand 
through   a   very  small  orifice    into   a 
chamber  open  to  the  atmosphere.    The 
excess  of  heat  liberated  serves  first  to 
evaporate    any  moisture  present   and 
then  to  superheat  the  steam  at  the 
lower   pressure.     From   the  observed 
temperature  and  pressures  it  is  easy   GRADUATED 
to  calculate,  with  the   aid   of    steam 
tables,  the  percentage  of  moisture  in 
the  original  sample. 

The  limit  of  the  throttle  calorimeter 
depends  upon  the  steam  pressure  and 
is  about  3  per  cent  of  moisture  at  80 
pounds  pressure  and  about  5  per  cent 
at  200  pounds.  For  steam  containing 
greater  percentages  of  moisture  the 
separating  calorimeter,  Fig.  434,  is 

sometimes  used.  This  instrument  is  virtually  a  steam  separator  and 
mechanically  separates  the  moisture  from  the  sample  of  steam. 


SEPARATOR 


STEAM  JACKET 


WATER 


DISCHARGE 
ORIFICE 


FIG.  434.    Carpenter  Separating 
Calorimeter. 


746 


STEAM  POWER  PLANT  ENGINEERING 


-    THERMOMtTCft 


TO  ATMOSPHERE 

FIG.  435.    A  Typical  Throttling  Calorimeter. 


THERMOMETER  WELL 


DRAIN  COCK 

FIG.  435a.    Ellison  Universal  Steam  Calorimeter. 


TESTING  AND  MEASURING  APPARATUS 


747 


The  water  thus  separated  collects  in  a  reservoir  provided  with  gauge 
glass  and  graduated  scale,  while  the  dry  steam  passes  through  an 
orifice  to  the  atmosphere.  The  weight  of  dry  steam  per  unit  of  time 
is  indicated  on  the  gauge,  calculated  according  to  Napier's  rule,  or  may 
be  determined  by  condensing  and  weighing.  The  accuracy  of  the 
moisture  determination  is  greatly  affected  by  the  difficulty  of  obtain- 
ing true  samples  of  steam  containing  large  percentages  of  moisture. 

Fig.  435a  shows  the  Ellison  universal  steam  calorimeter,  which 
combines  the  superheating  and  throttling  principles  and  is  adapted  to 
steam  of  any  degree  of  wetness.  The  separating  chamber  is  provided 
with  a  gauge  glass,  not  shown,  for  indicating  the  weight  of  water  which 
accumulates  only  when  the  steam  is  too  wet  to  be  superheated. 

Throttling  Calorimeters:  Power,  Dec.,  1907,  p.  891;  Trans.  A.S.M.E.,  17-151, 
175,  16-448;  Engr.  U.S.,  Feb.  15,  1907,  p.  219. 

Separating  Calorimeters:  Trans.  A.S.M.E.,  17-608;  Engr.  U.S.,  Feb.  15,  1907, 
p.  219. 

Universal  Calorimeter:   Trans.  A.S.M.E.,  11-790. 

Thomas  Electrical  Calorimeter:   Power,  Nov.,  1907,  p.  791. 

411.  Fuel  Calorimeters.  —  The  analysis  and  heat  evaluation  of 
fuel  require  considerable  time  and  skill  and  much  costly  apparatus, 


INSULATION 
BOMB 

PLATINUM    PAN 
WATER 

ELECTRODE 
IGNITION    WIRE 
STIRRING   DEVICE 


FIG.  436.    Mahler  Bomb  Calorimeter. 


hence  in  most  power  plants  it  is  customary  to  depend  upon  a  specialist 
to  whom  samples  are  submitted  from  time  to  time.     In  many  large 


748 


STEAM  POWER  PLANT  ENGINEERING 


stations,  however,  the  conditions  often  warrant  the  establishment  of  a 
testing  laboratory  equipped  for  the  proximate  analysis  of  coal  and  the 
determination  of  the  calorific  value  of  the  solid,  liquid,  or  gaseous  fuel 
used.  The  Mahler  bomb  calorimeter  illustrated  in  Fig.  436  is  the 
most  accurate  and  satisfactory  device  for  solid  and  liquid  fuels  but  is 
comparatively  expensive.  The  instrument  consists  of  a  steel  shell  or 
"  bomb  "  of  great  strength,  lined  with  porcelain  or  platinum,  into 
which  a  weighed  sample  of  the  fuel  is  introduced  and  burned  on  a 
platinum  pan  in  the  presence  of  oxygen  under  a  pressure  of  about  300 
pounds  per  square  inch.  The  charge  is  ignited  by  an  electric  current. 
During  combustion  the  bomb  is  submerged  in  a  known  weight  of 
water  which  is  kept  constantly  agitated.  The  calorific  value  is  calcu- 
lated from  the  observed  rise  in  temperature  due  to  the  heat  evolved, 
proper  corrections  being  made  for  the  water  equivalent  of  bomb  and 
appurtenances,  heat  given  up  by  the  igniting  current,  and  for  radiation 
or  absorption  of  heat  from  the  surrounding  air. 


COMPRESSED 
FIBER 


STIRRER 


FIG.  437.    Parr  Fuel  Calorimeter. 


The  Parr  calorimeter,  Fig.  437,  is  an  inexpensive  instrument,  very 
simple  in  operation,  and  gives  results  which  are  sufficiently  accurate 
for  all  practical  purposes.  The  weighed  sample  of  coal,  together  with 
a  quantity  of  sodium  peroxide  which  supplies  the  oxygen  for  com- 
bustion, is  introduced  into  the  cartridge.  Means  are  provided  for 


TESTING  AND  MEASURING  APPARATUS  749 

rotating  the  cartridge  when  submerged  in  the  calorimeter,  the 
attached  vanes  agitating  the  water  to  maintain  uniform  temperature. 
The  charge  is  fired  either  electrically  or  by  introducing  a  short  piece  of 
hot  wire  through  the  conical  valve.  The  calorific  value  is  calculated 
from  the  observed  rise  in  temperature  and  the  constants  of  the  instru- 
ment. Among  other  forms  of  instruments,  in  more  or  less  general  use 
and  which  give  very  satisfactory  results,  may  be  mentioned  the  Car- 
penter calorimeter  and  the  Thompson  calorimeter. 

Mahler  Bomb  Calorimeter:   Engr.  U.S.,  Jan.  1,  1907,  p.  68. 

Parr  Fuel  Calorimeter:  Power,  July,  1907,  p.  499;  Engr.  U.S.,  April  1,  1903; 
Jour.  Am.  Chem.  Soc.,  22-246;  The  Calorific  Value  of  Fuels,  Poole;  Gas  and  Fuel 
Analysis,  Gill;  Eng.  Chem.,  Stillman;  Chem.  Technology,  Groves  and  Trop. 

Carpenter  Coal  Calorimeter:   Trans.  A.S.M.E.,  16-1040. 

Thompson  Coal  Calorimeter:   Jour.  Soc.  Chem.  Ind.  (1906),  25-409. 

Junkers  Gas  Calorimeter:  Jour.  Soc.  Chem.  Ind.  (1895),  14-631 ;  Stevens  Indicator, 
Jan.,  1905,  p.  31. 

Comparison  of  Different  Types  of  Calorimeters:  Jour.  Soc.  Chem.  Ind.  (1903), 
22-1230. 

41  la.  Hamler-Eddy  Smoke  Recorder.  —  This  apparatus  consists 
essentially  of  a  small  motor-driven  vacuum  pump,  which  draws  a  con- 
tinuous sample  of  the  products  of  combustion  from  the  uptake  breeching 
or  stack  and  discharges  it  against  a  paper-covered  drum  revolved  by 
clockwork.  The  density  of  the  smoke,  the  time  at  which  visible  smoke 
is  being  emitted  and  the  duration  of  the  smoke  production  period  are 
automatically  recorded  on  the  paper  by  the  smoke  itself.  Before 
reaching  the  pumps  the  gases  pass  through  a  glass  "  emergency  "  con- 
denser and  a  large  portion  of  the  vapor  content  is  removed.  The  pump 
discharges  the  partially  dried  gases  against  a  surface  of  sulphuric  acid 
(which  removes  the  last  trace  of  moisture)  and  forces  the  smoke  in  the 
form  of  a  small  jet  of  dry  powder  onto  the  surface  of  the  recording  paper. 
The  sampling  tube  leading  from  the  flue  to  the  pump  is  connected  with 
a  steam  line  and  is  "  blown  out  "  each  time  a  card  is  changed.  The 
instrument  is  very  compact  and  portable  and  may  be  placed  anywhere 
with  respect  to  the  chimney.  A  number  of  these  appliances  in  Chicago 
power  plants  are  giving  excellent  satisfaction. 


CHAPTER  XIX. 


TYPICAL  SPECIFICATIONS. 

412.  Sample  Specifications  for  a  Cross  Compound  Non-Condensing 
Engine. — For  and  in  consideration  of  the  amount  and  terms  named  in  the  letter 
accompanying  this  specification,  and  of  the  same  date,  we  propose  to  furnish 
f.o.b.  cars  at  our  factory,  Elizabethport,  N.J.,  for  account  of  THE  ARMOUR 
INSTITUTE  OF  TECHNOLOGY,  Chicago,  111.,  ONE  BALL  &  WOOD  HORIZONTAL 
FOUR- VALVE  (CORLISS)  CENTER-CRANK  ENGINE,  designed  for  direct  connection 
to  a  direct-current  generator,  as  follows : 

General  Horse  power 350  to  375 

Dimensions.  Diameter  of  cylinders H.P.,  17;  L.P.,  27  inches 

Length  of  stroke 18  inches 

Revolutions  per  minute 175  to  200 

Governor  wheel diameter,  90;  face,  21  inches 

Width  of  belt  (if  belted)    20  inches 

Diameter  of  steam  pipe 7  inches 

Diameter  of  exhaust  pipe 10  inches 

Crosshead  pins 6  in.  long,  8  in.  diameter 

Crank  pins 9  in.  long,  9£  in.  diameter 

Main  bearings 20  in.  long,  9  in.  diameter 

Wearing  surface  of  crossheads 242  square  inches 

Width  of  engine  over  all 14  feet  6  inches 

Length  of  engine  over  all 20  feet  3  inches 

Weight  complete 60,000  pounds 

Rating-.  The  rated  power  of  the  engine  specified  is  based  on  an  initial 

pressure  of  120  pounds  (measured  in  the  cylinder),  cutting  off  at 
about  one  third  stroke  in  both  cylinders,  without  vacuum,  when 
operating  at  200  revolutions  per  minute. 

Fittings.  With  each  engine  is  furnished  the    following  complete  list  of 

fittings : 

One  extended  shaft  (omitted  if  belted  engine), 

One  sub-base  with  extension  for  dynamo  (omitted  if  belted 

engine), 

One  self-oiling  outboard  bearing  (omitted  if  belted  engine), 
One  throttle  valve, 
One  lubricating  system  consisting  of  pipes,  sight  feeds,  and 

oil  reservoir, 

One  cylinder  lubricator,  nickel  plated, 
750 


TYPICAL  SPECIFICATIONS  751 

Fittings  —  Continued. 

One  set  special  steel  wrenches, 

One  socket  wrench  for  piston, 

One  socket  wrench  for  connecting  rod  bolts, 

One  steel  wrench  for  hexagon  nuts, 

One  connecting  rod  set  screw  wrench, 

One  monkey  wrench, 

One  spanner  for  valve  stem  gland, 

One  eye  bolt  for  pillow  block  cap, 

Two  push-off  bolts, 

One  set  grease  cups, 

Two  oil  cups, 

One  hand  oil  pump, 

One  set  brass  oil  cups, 

Three  nipples  for  drip  pipes,  for  frame, 

Four  nipples  for  cylinder  drips,  3  inches  long, 

One  nipple  for  throttle  bleeder, 

One  globe  valve  for  throttle  bleeder, 

Four  globe  valves  for  cylinder  drips, 

One  set  foundation  bolts,  nuts,  and  plates, 

One  template  for  locating  bolts, 

One  governor  wheel  and  keys, 

One  balance  wheel  and  keys  (omitted  if  direct  connected 
engine), 

Packing  for  piston  and  valve  rods, 

One  one-gallon  can  cylinder  oil, 

One  one-gallon  can  engine  oil, 

Two  cans  grease, 

Two  wedges  for  wheels, 

Cylinders.  Cylinders  are  made  of  hard  and  close-grained  iron,  and  under 
the  influence  of  oil  and  wear  the  walls  will  rapidly  acquire  a  fine, 
smooth  glaze.  Radiation  is  prevented  by  a  thick  jacket  of 
asbestos  cement,  outside  of  which  is  neatly  fitted  an  orna- 
mented jacket.  Openings  and  globe  valves  are  provided  for 
drainage. 

Connecting        The   connecting  rods  are  of   forged  machinery  steel   of  low 
Rods.  carbon  and  fitted  with  heavy  straps  with  keys  and  bolts  for 

adjustment. 

Crank  Pin          The  crank  pin  boxes  are  of    cast  iron  lined    with    Babbitt 
Boxes.  metal. 

Crosshead          The  crosshead  boxes  are  of  the  best  quality    of    phosphor 
Boxes.  bronze. 

Crossheads.  The  crossheads  are  of  cast  iron  faced  with  Babbitt  metal  on  the 
wearing  surfaces.  The  crosshead  pins  are  pressed  into  the  cross- 
heads. 


752 


STEAM  POWER  PLANT  ENGINEERING 


Piston. 


Valves. 


Speed. 


Frame. 


Main 
Bearings. 


Guides, 


The  piston  rods  are  of  special  hammered  steel,  threaded  and 
screwed  into  the  crossheads,  and  locked  fast  with  special  nut 
counterbored  at  the  end  to  cover  threads,  finished  and  case- 
hardened.  The  other  ends  of  the  rods  will  be  fitted  to  the  pistons 
with  thread  and  locked  with  nut.  The  pistons  will  be  fitted  with 
two  rings  turned  eccentric  and  cut  open  at  the  thinnest  part,  the 
ends  being  halved  so  as  to  lap  when  in  position. 

Both  the  admission  and  exhaust  valves  are  of  the  Corliss 
pattern.  The  former  are  provided  with  double  ports,  and  are 
actuated  from  a  wrist  plate  receiving  its  motion  from  the  governor 
placed  in  the  fly  wheel  of  the  engine.  This  governor  controls  the 
valves  of  both  the  high  and  low-pressure  cylinders  and  possesses 
a  range  of  cut-off  from  0  to  about  f  stroke. 

The  exhaust  valves  are  driven  from  a  wrist  plate  through  an 
adjustable  eccentric  by  which  any  desired  degree  of  compression 
can  be  obtained. 


The  use  of  Corliss  valves,  arranged  as  described  in  the  foregoing 
paragraph,  permits  an  increased  speed  over  the  common  type 
of  Corliss  engine  with  releasing  gear,  and  while  yielding  the 
same  economy  dispenses  with  many  working  parts,  and,  what 
is  more  important,  with  the  large  and  cumbersome  fly 
wheel  which  has  so  often  proved  a  source  of  danger  in  slow- 
speed  engines. 

The  frame  is  proportioned  for  great  strength  and  the  metal  is 
placed  where  it  is  most  needed.  An  oil  groove  is  cast  around  the 
bottom  of  the  frame  to  protect  the  foundation. 

Main  bearings  are  fitted  with  removable  Babbitt  shells  which 
can  be  replaced  when  necessary.  Special  care  is  taken  to  have 
these  bearings  of  ample  length  to  support  the  wheels  and  stand 
the  strain  of  power  transmission  or  the  weight  of  armature  when 
direct  connected.  In  the  latter  case  a  self-oiling  outboard  bearing 
is  provided  to  carry  the  outer  end  of  shaft. 

The  guides  are  known  as  the  locomotive  pattern  and  are  inter- 
changeable. They  are  carefully  scraped  to  surface  plates  and 
provision  made  for  taking  up  wear. 

Crank  Shaft.  The  crank  shaft  is  of  the  best  quality  of  steel,  being  carefully 
counterbalanced  by  cast-iron  disks  in  which  the  necessary  weight 
is  placed.  In  direct  connected  engines  this  shaft  is  either 
extended  in  one  piece  to  carry  the  armature  or  made  in  two 
pieces  and  coupled. 

The  governor  is  of  the  inertia  type  and  has  a  swinging  eccen- 
tric, the  eccentric  center  moving  across  the  end  of  the  shaft  about 
an  outside  point,  and  giving  a  lead  which  varies  with  the  point  of 


Governor. 


TYPICAL  SPECIFICATIONS 


753 


Governor  —  Continued. 

cut-off  from  a  maximum,  at  the  latest  point,  to  zero,  when  the 
governor  weights  occupy  their  extreme  outward  position.  Alter- 
ation in  speed  is  obtained  by  changing  the  amount  of  weight  in 
the  pockets  of  the  lever  arm. 

Balance.  The  balance  wheel  (in  the  case  of  belted  engine)  is  made  with 

a  flanged  rim  and  with  a  split  hub,  the  hub  being  secured  to  the 
shaft  by  a  bolt.  The  other  keys  are  square,  with  parallel  sides, 
and  are  inserted  without  driving. 

Oiling.  The  oiling  system  consists  of  a  simple  oil  reservoir  which  sup- 

plies oil  through  a  system  of  pipes  to  the  points  of  the  engine 
needing  lubrication.  After  fulfilling  its  functions  this  oil  is  drained 
and  can  be  used  anew.  This  does  away  with  the  old  cumbersome 
oil-cup  system  and  has  the  great  advantage  of  deli vering  clean  oil 
to  the  engine. 

Guarantees.  Material  We  guarantee  that  the  material  and  workmanship 
are  of  the  best  and  that  all  working  parts  having  flat  surfaces  are 
scraped  to  surface  plates. 

Regulation.  That  the  engine  shall  regulate  within  2  per  cent 
under  changes  of  load  within  the  range  of  the  governor,  and  that 
no  reduction  of  boiler  pressure  shall  reduce  the  speed  until  the 
latest  point  of  out-off  is  reached. 

Steam  Consumption.  That  the  steam  consumption,  when  the 
engine  is  developing  its  rated  power  at  125  pounds  pressure 
and  no  vacuum,  shall  not  exceed  22  pounds  of  dry  steam  per 
indicated  horse  power  per  hour;  that  the  clearance  shall  not 
exceed  8  per  cent. 

Drawings.  With  the  engine  is  furnished  a  drawing  showing  its  details, 
together  with  foundation  plans. 

Preparation       Every  engine  is  completely  erected  at  our  works  before  ship- 

f or  Ship-         ment.     The  castings  are  rubbed  smooth,  carefully  filled,  and  the 

ment.  engine  given  two  good  coats  of  standard  shop  color.     All  bright 

parts  are  carefully  protected  against  corrosion.     The  engine  is 

dismantled,  the  small  parts  being  boxed,  and  in  the  case  of  export 

shipment  the  larger  pieces  crated. 

Erection.  Full  drawings  and  directions  for  erecting  the  engine  will  be 

furnished.  Template,  foundation  bolts,  nuts,  and  plates  to  be 
shipped  in  advance  if  necessary,  and  by  freight. unless  otherwise 
directed. 

If  requested  we  will  furnish  the  services  of  an  expert  to  superin- 
tend the  erection  of  this  engine  at  the  rate  of  $5  per  day  added 
to  his  traveling  expenses  and  board,  the  purchaser  to  furnish  all 
laboring  help. 


754  STEAM  POWER  PLANT  ENGINEERING 

Terms.  One-half  cash  on  presentation  of  bill  of  lading,  balance  on 

completion  of  erection. 

The  title  to  the  apparatus  herein  sold  shall  not  pass  from  The 
Ball  and  Wood  Company  until  all  payments  hereunder  (including 
deferred  payments,  if  any)  shall  have  been  fully  made  in  cash. 
The  purchaser  agrees  to  do  all  acts  necessary  to  perfect  and  main- 
tain such  retention  of  title  in  the  said  Company.  All  previous 
communications  between  the  parties  hereto,  verbal  or  written,  are 
hereby  abrogated  and  withdrawn,  and  this  proposal,  when  duly 
signed  and  approved,  constitutes  the  agreement  between  the 
parties  hereto,  and  no  modification  of  this  accepted  agreement 
shall  be  binding  upon  the  parties  hereto  or  either  of  them  unless 
such  modification  shall  be  in  writing,  duly  accepted  by  the  pur- 
chaser and  approved  by  an  executive  of  the  Company. 

Limit.  Prices  subject  to  revision  after  thirty  days.     Delivery  subject 

to  strikes,  accidents,  or  causes  beyond  our  control. 

413.   Specifications    for   Horizontal   Tubular    Steam   Boiler.*  —  The 

following  specifications  for  one  54-inch  horizontal  return  tubular  steam 
boiler,  pressure  125  pounds,  were  prepared  by  the  HARTFORD  STEAM 
BOILER  INSPECTION  AND  INSURANCE  COMPANY  for  the  ARMOUR  INSTI- 
TUTE OF  TECHNOLOGY,  Chicago. 

The  boiler  to  conform  to  the  following  conditions  and  requirements : 

Type  and  It  is  to  be  of  the  horizontal  tubular  type,  set  with  overhanging 

General          front,  and  all  parts  and  pieces  are  to  be  designed  accordingly. 
Dimensions.       It  is  to  be  17  feet  2  inches  long,  outside,  and  54  inches  in  dia- 
meter, measured  on  the  outside  of  the  smallest  ring  of  plates. 
Heads  are  to  be  16  feet  0  inches  apart,  outside. 

Materials:          Shell  plates  are  to  be  three-eighths  of  an  inch  thick  on  the 

Quality,  edges,  of  open-hearth  fire-box  steel,  having  a  tensile  strength  of 

Thickness,      not  less  than  55,000  pounds  nor  more  than  62,000  pounds  per 

and  Tests.      square  inch  of  section,  and  an  elastic  limit  of  not  less  than  half 

the  tensile  strength,  with  not  less  than  56  per  cent  of  ductility,  as 

indicated  by  contraction  of  area  at  point  of  fracture  under  test, 

and  by  an  elongation  of  25  per  cent  in  a  length  of  8  inches. 

Heads  are  to  be  one-half  of  an  inch  thick,  of  best  open-hearth 
flange  steel.  All  plates,  both  of  shell  and  heads,  are  to  be  plainly 
stamped  with  name  of  maker,  brand,  and  tensile  strength ;  brands 
so  located  that  they  may  be  seen  on  each  plate  after  the  boiler  is 
finished. 

Each  shell  plate  is  to  bear  a  coupon  which  shall  be  sheared  off, 
finished  up,  and  tested  by,  or  for,  the  maker  of  the  boiler,  at  his 
expense.  Each  coupon  is  to  fulfill  the  foregoing  requirements  as 

*  Drawings  have  been  omitted. 


TYPICAL  SPECIFICATIONS 


755 


Materials:  Quality,  Thickness,  and  Tests—  Continued. 

to  strength  and  ductility,  and  stand  bending  down  double  when 
cold,  when  red  hot,  and  after  being  heated  and  quenched  in  cold 
water,  without  signs  of  fracture.  There  is  not  to  be  more  than 
0.035  per  cent  of  sulphur,  nor  more  than  0.035  per  cent  of  phos- 
phorus in  the  chemical  composition  of  the  plates  and  heads.  All 
plates  failing  to  pass  these  tests  will  be  rejected.  All  tests  and 
inspections  of  material  may  be  made  at  the  place  of  manufacture 
prior  to  shipment.  Certified  copies  of  report  of  tests  must  be 
sent  to  the  Hartford  Steam  Boiler  Inspection  and  Insurance 
Company,  Hartford,  Conn. 

Riveting.  The  longitudinal  seams  are  to  be  of  the  double-riveted  butt- 

joint  type  with  double  covering  strips.  They  are  to  be  arranged 
to  come  well  above  the  fire  line  of  the  boiler,  and  break  joints 
in  the  3  ring  courses  in  the  usual  manner.  The  plates  are  to  be 
planed  on  the  caulking  edges  before  rolling. 

All  dimensions  and  proportions  are  to  be  shown  on  accompany- 
ing drawing  No.  1502. 

The  girth  seams  are  to  be  of  the  single-riveted  lap-joint  type; 
rivets  to  be  of  same  size  as  those  in  longitudinal  seams,  and 
pitched  2£  inches  apart  from  center  to  center;  the  distance  from 
center  of  rivet  to  the  edge  of  the  plate  to  be  equal  to  1^  times 
the  diameter  of  rivet  hole. 

The  rivet  holes  are  to  be  either  drilled  in  place,  or  punched  at 
least  one-quarter  of  an  inch  less  than  full  size ;  if  the  latter  method 
is  used,  the  plates,  after  punching,  are  to  be  rolled  and  bolted 
together,  and  the  rivet  holes  drilled  in  place  one-sixteenth  of  an 
inch  larger  than  the  diameter  of  the  rivets.  The  plates  are  then 
to  be  disconnected.  All  burrs  are  to  be  removed  from  the  edges 
of  the  holes.  Should  any  holes  be  in  the  least  out  of  true,  they 
are  to  be  brought  in  line  with  a  reamer  or  drill;  if  a  drift-pin  is 
used  for  this  purpose  the  boiler  will  be  rejected. 

All  rivets  are  to  be  driven  by  hydraulic  pressure,  wherever 
possible,  and  allowed  to  cool  and  shrink  under  pressure.  This  pres- 
sure is  to  completely  fill  the  rivet  holes,  producing  a  tight  joint. 

Rivet  Ham-  The  rivets  are  to  be  of  the  best  quality  of  iron  or  soft  steel, 
mer  Tests,  capable  of  being  hammered  flat,  when  cold,  to  a  thickness  of 
one-half  their  original  diameter,  or  when  hot,  to  one-third  their 
original  diameter,  without  showing  signs  of  fracture.  In  the 
absence  of  physical  test,  it  is  understood  that  the  contractor 
guarantees  the  above  quality  of  rivets. 

Braces,  There  are  to  be  20  braces  1^  inches  in  diameter  in  the  boiler, 

10  above  the  tubes  on  front  head,  and  10  on  rear  head,  of  the 
crow-foot  form,  arranged  as  shown  on  drawing.  None  of  them 
is  to  be  less  than  3  feet  6  inches  long,  and  each  is  to  be  fastened 


756  STEAM  POWER  PLANT  ENGINEERING 

Braces  —  Continued. 

to  shell  and  heads  by  two  seven-eighths  inch  rivets  at  each  end; 
or  solid  steel,  diagonal  braces  of  approved  pattern,  and  of  equal 
strength  to  the  former,  may  be  used.  Care  is  to  be  exercised  in  set- 
ting them  that  they  may  bear  uniform  tension.  Crow-foot  braces 
may  be  flat  in  body,  if  of  equal  strength  to  those  specified  above. 
Braces  There  are  to  be  4  braces  below  the  tubes  in  the  boiler.  Two 

below  of  these  are  to  be  through  braces  extending  from  head  to  head. 

Tubes.  Each  brace  is  to  be  1^  inches  in  diameter,  with  a  fork  formed  on 

rear  and  secured  with  a  1^-inch  turned  bolt  and  nut  to  a  crow- 
foot securely  riveted  to  rear  head;  these  are  the  inner  or  central 
braces.  The  front  end  of  brace  is  to  be  upset  to  a  diameter  of  If 
inches,  threaded  and  secured  to  front  head  with  a  nut  and  washer 
on  both  the  inside  and  outside  of  head. 

The  2  remaining  braces  are  each  to  be  1^  inches  in  diameter, 
and  secured  to  rear  head  in  same  manner  as  the  through  braces; 
the  front  end  of  the  brace  is  to  be  extended  forward,  fitted  to  side 
of  shell,  and  riveted  there  with  two  1-inch  rivets.  All  to  be 
substantially  as  shown  on  accompanying  diagram  of  tube  head 
No.  2431. 

Tubes:  Size,  There  are  to  be  36  lap-welded  or  seamless-drawn  tubes,  of  the 
Number,  and  best  quality  with  regard  to  tensile  strength  and  ductility.  They 
Arrange-  are  to  be  round,  straight,  and  free  from  all  surface  defects,  prop- 
ment.  erly  annealed  on  their  ends,  and  guaranteed  by  the  manufac- 

turers to  have  been  tested  to  at  least  five  hundred  (500)  pounds 
per  square  inch  internal  hydrostatic  pressure.  Each  tube  is  to 
be  4  inches  in  diameter,  16  feet  0  inches  long,  and  not  less  than 
standard  thickness,  set  in  vertical  rows,  with  a  clear  space 
between  them,  vertically  and  horizontally,  of  1  inch,  except  the 
central  vertical  space,  which  is  to  be  2  inches,  as  shown  on  accom- 
panying diagram  of  tube  head  No.  2431. 

Holes  for  tubes  are  to  be  neatly  chamfered  off  on  outside. 
Tubes  to  be  set  with  a  Dudgeon  expander,  and  beaded  down  at 
each  end.  Tube  holes  may  be  drilled  and  reamed,  or  may  be 
punched  one-quarter  inch  less  than  full  size,  then  rose  bitted  to 
exact  diameter. 

Manholes.  There  are  to  be  two  manholes,  one  11  x  15  inches,  with  pressed 

steel  frame,  double  riveted  to  inside  of  shell  on  top,  and  one 
10  x  15  inches,  flanged  in  front  head  below  tubes,  with  suitable 
plates,  yokes,  and  bolts,  the  proportions  of  the  whole  such  as  will 
make  them  as  strong  as  any  portion  of  the  shell  of  like  area. 

Boiler  The  boiler  to  be  suspended  from  steel  I  beams,  6  inches  deep, 

Supports.        12J  pounds  per  foot,  by  means  of  eye  or  U  bolts  and  plate  loops. 

There  are  to  be  6  loops,  2  on  each  side  of  the  boiler,  securely 

riveted  to  boiler  shell.     The  I  beams  are  to  be  supported  on  cast- 


TYPICAL  SPECIFICATIONS 


757 


Boiler  Supports  —  Continued. 

iron  columns  of  square  or  rectangular  section  6  inches  square, 
three-quarters  inch  thick.  Each  pair  of  beams  is  to  be  connected 
together,  3  inches  apart,  by  tie-bolts  and  cast-iron  separators ;  one 
separator  near  each  end,  and  others  at  intervals  of  about  five 
feet.  The  top  and  bottom  flanges  of  columns  are  to  be  faced  true. 
The  whole  system  of  suspension  is  to  be  made  in  the  best  man- 
ner, properly  arranged  to  allow  free  expansion  of  the  boiler, 
securely  held  and  supported  in  every  direction,  amply  strong  in 
every  part,  and  finished  complete. 

Nozzles.  There  are  to  be  two  heavy  cast  nozzles,  made  of  gun-iron  or 

steel,  one  4  inches  internal  diameter  for  steam-pipe  connection, 
and  one  6  inches  internal  diameter  for  safety-valve  connection, 
each  accurately  squared  on  top  flange,  and  securely  riveted  to 
boiler  on  top.  Forged  or  pressed  steel  pipe  flanges  may  be  used 
in  place  of  nozzles. 

The  flanges  of  the  nozzles  to  correspond  in  diameter  and  thick- 
ness with  standard  extra  heavy  pipe  fittings. 

Smoke  There  is  to  be  an  opening  10  by  62  inches  cut  out  of  front 

Opening.         connection  on  top  for  attachment  of  uptake  or  flue. 

Feed  Pipe.  There  is  to  be  a  hole  tapped  in  front  head  for  a  brass  bushing, 
3  inches  above  the  top  of  upper  row  of  tubes,  and  16  inches  from 
center  of  boiler,  on  left-hand  side,  for  li-inch  feed-pipe  connection. 
The  bushing  is  to  be  not  less  than  2  inches  long,  to  permit  both 
the  external  and  internal  feed  pipes  to  be  screwed  into  it  not 
less  than  seven-eighths  inch. 

Also  furnish  and  put  in  a  1^-inch  feed  pipe  extending  from 
front  head  back  to  within  two  feet  of  rear  head  of  boiler,  thence 
across  the  boiler  to  near  shell  on  right-hand  side.  On  this  end 
place  an  elbow  with  the  outlet  pointed  down  as  shown  on  draw- 
ings. Feed  pipe  is  to  be  properly  hung  from  the  braces. 

Blow-off  There  is  to  be  an  extra  heavy  pressed  steel  pipe  flange,  riveted 

Pipe  Con-       to  bottom  of  shell,  near  rear  end,  and  tapped  to  receive  a  4-inch 
nection.          extra  heavy  blow-off  pipe.     Blow-off  valve  and  fittings  to  be 

extra  heavy. 

Fusible  There  is  to  be  a  fusible  plug  in  rear  head,  two  inches  above 

Plug.  top  of  upper  row  of  tubes. 

Fittings.  There  is  to  be  furnished  one  pop  safety  valve  3  inches  .in 

diameter,  one  6-inch  steam  gauge,  three  three-quarters-inch  gauge 
cocks,  and  one  three-quarters-inch  gauge  glass  12  inches  long,  all 
to  be  of  approved  pattern,  and  the  necessary  holes  to  be  made  for 
their  proper  connection.  If  combination  water  column  is  used, 
the  steam  and  water  connections  between  it  and  the  boiler  must 
be  made  by  pipe  not  less  that  l£  inches  in  diameter. 


758 


STEAM  POWER  PLANT  ENGINEERING 


Castings  There  is  to  be  furnished  a  substantial  cast-iron  front,  with  all 

for  Setting,  necessary  anchor  bolts,  10  feet  long,  closely  fitting  front  connec- 
tion doors  with  suitable  fastening  to  prevent  warping,  closely 
fitting  furnace  doors  with  liner  plates,  rear  connection  door 
16  x  24  inches,  with  liner  plates,  grate  bars  for  grate,  pattern  to 
be  selected  by  purchaser  of  boiler,  54  inches  long  by  48  inches  wide, 
with  suitable  bearer  bars  for  same,  arch  bars  for  rear  connection, 
and  all  buckstaves,  with  the  necessary  bolts  or  tie  rods,  and  all 
other  castings  or  ironwork  of  any  description  necessary  for  the 
proper  construction  and  setting  of  the  boiler  complete. 

In  General.  The  intent  of  the  foregoing  specification  is  to  provide  for 
material  and  workmanship  of  the  best  quality,  and  any  details  of 
equipment  not  mentioned  in  this  specification,  or  not  shown  on  the 
drawings,  but  necessary  for  the  proper  completion  of  the  boiler 
ready  for  operation,  and  to  be  hereafter  contracted  for,  must  be 
of  equally  good  quality. 

The  size  and  description  of  parts  are  to  conform  substantially 
to  the  details  of  the  accompanying  plan,  and  the  boiler,  complete, 
is  to  be  delivered  at and  all  of  the  material  and  workman- 
ship is  to  be  subjected  to  the  inspection  and  approval  of  the 
Hartford  Steam  Boiler  Inspection  and  Insurance  Company. 

414.   Specifications    for    Barometric    Condenser    and    Auxiliaries.  — 

The  following  sample  specifications  for  a  barometric  condenser  will  give 
some  idea  of  the  various  items  called  for  in  the  purchase  of  a  condenser 
and  appurtenances,  the  italicized  items  being  specified  by  the  purchaser. 

"  We    submit    herewith  our  tender  for  one    condensing 

plant  as  follows: 

Rated  The  condenser  and  auxiliary  machinery  will  have  sufficient 

Capacity.  capacity  to  condense  250  pounds  of  steam  per  minute  (equivalent 
to  the  steam  exhausted  from  engines  developing  1000  horse 
power  on  a  basis  of  15  pounds  of  steam  per  horse  power  per  hour) 
when  supplied  with  cooling  water  at  a  temperature  of  70  degrees 
Fahrenheit,  and  maintaining  a  vacuum  at  the  condenser  of  26 
inches  of  mercury. 

Capacity  un-      The  plant  will  also  have  to  condense  the  quantities  of  steam 
der  Variable  under  the  varying  conditions  as  stated  below : 
Conditions. 


Steam  Condensed  per 
Minute,  Pounds. 

Temperature  of  Cooling 
Water,  Degrees  F. 

Vacuum  Maintained  at 
Condenser,  Inches  of 
Mercury. 

250 
300 
340 
250 
290 

70 
70 
70 
80 
80 

26 
25 
24 
25 
24 

TYPICAL  SPECIFICATIONS 


759 


Quantity  of 

Cooling 

Water. 

Apparatus 

Furnished. 


Price  and 
Delivery. 

Terms. 


Superin- 
tendence. 

Steam 
Pressure. 


Head 

Pumped 

against. 

Power 
Consump- 
tion of 
Auxiliaries. 


The  volume  of  cooling  water  required  when  the  condenser  is 
working  under  the  above  conditions  will  be  from  550  gallons  to 
650  gallons  per  minute. 

The  apparatus  to  be  furnished  by  us  will  consist  of: 

One  cast-iron  condensing  vessel,  complete  with  barometric 
tubes  and  foot  valves. 

One  automatic  vacuum  regulator. 

A  structural  steel  framework  for  supporting  the  condensing 
vessel. 

One  positive  rotary  pump  for  supplying  the  cooling  water  to 
the  condenser. 

One  "dry"  air  pump. 

One  horizontal  steam  engine,  arranged  to  drive  the  water  pump 
by  belt  and  the  air  pump  direct,  the  latter  placed  tandem  to  the 
engine ;  the  engine  is  to  be  fitted  with  a  suitable  governor  arranged 
for  variable  speeds.  Purchaser  to  furnish  belt. 

Two  pulleys,  one  for  the  engine  and  one  for  the  water  pump. 

One  vacuum  gauge. 

Four  thermometers. 

We  do  not  include  any  steam,  air,  or  water  pipes,  valves  nor 
foundation  bolts,  but  will  furnish  plans  showing  suitable  founda- 
tions and  general  arrangement  of  the  machinery. 

Our  price  for  one  barometric  condensing  plant,  as  described, 
including  all  royalties,  and  delivered  f.o.b.  cars  our  works,  is 


Payments  as  follows:  Monthly  payments  as  the  work  pro- 
gresses in  our  shops,  less  10  per  cent.  The  retained  percentage 
to  be  paid  when  the  condenser  is  started  in  service,  provided 
this  is  done  within  a  reasonable  time  after  completion. 

If  desired,  we  will  furnish  a  competent  machinist  to  superintend 
the  erection  and  starting  of  the  plant,  charging  extra  for  his  serv- 
ices, 50  cents  per  hour  and  his  traveling  and  boarding  expenses. 

The  engine  driving  the  air  and  water  pumps  will  be  capable  of 
starting  and  operating  the  plant  with  1 00  pounds  minimum  steam 
pressure,  and  will  be  built  strong  enough  to  work  under  135 
pounds  maximum  steam  pressure. 

The  water  pump  and  engine  driving  same  will  be  designed  to 
raise  the  injection  water  from  the  cold  well  to  the  condenser,  the 
level  of  the  water  in  the  cold  well  to  be  not  more  than  10  feet  below 
the  level  of  the  water  in  the  hot  well. 

The  engine  driving  the  air  and  water  pumps  will  require  approxi- 
mately 2  per  cent  of  the  main  engine  steam  when  operating  under 
rated  load  and  with  conditions  as  above  stated. 

The  rotary  pump  has  a  positive  displacement  (Bibus*  patent) 


760  STEAM  POWER  PLANT  ENGINEERING 

Power  Consumption  of  Auxiliaries  — Continued. 

and  is  of  substantial  construction.  All  the  power  for  propelling 
the  water  is  transmitted  through  the  main  shaft,  the  office  of  the 
gears  being  simply  to  keep  the  sealing  runner  in  time  with  the 
propelling  runner.  Stuffing  boxes  are  placed  between  the  pump 
chamber  and  bearings  to  prevent  any  grit  in  the  water  coming 
into  contact  with  the  journals. 

The  engine  and  air  pump  is  of  the  crank  and  fly  wheel  type, 
the  engine  being  fitted  with  a  suitable  governor  arranged  for  vari- 
able speeds,  and  the  air  pump  with  the patented  slide  valve. 

415.  Specification  for  Steam,  Exhaust,  Water,  and  Condenser  Piping 
for  an  Electric  Power  Station.*  —  It  is  intended  that  this  specification 
shall  cover  the  complete  installation  of  steam  piping,  exhaust  piping, 
injection  and  discharge  piping,  drain,  drip,  blow-off,  and  boiler-feed 
piping,  water  piping,  valves,  separators,  anchors,  fittings,  etc.,  sub- 
stantially as  shown  on  the  accompanying  drawing  or  hereinafter 
described. 

All  the  materials  used  throughout  must  be  the  best  of  their  respec- 
tive kinds,  subject  to  the  inspection  and  approval  of  the  engineer  of 
the  purchaser. 

The  entire  work  provided  for  in  this  specification  is  to  be  constructed 
and  finished  in  every  part  in  a  good,  substantial,  and  workmanlike 
manner,  according  to  the  accompanying  drawings  and  this  specifica- 
tion to  the  full  intent  and  meaning  of  the  same,  and  everything  necessary 
for  the  proper  and  complete  execution  of  the  plans  and  drawings, 
whether  the  same  may  have  been  herein  particularly  specified  or  not, 
or  indicated  in  the  plans  referred  to,  to  be  done  and  furnished  in  a 
manner  corresponding  with  the  rest  of  the  work  as  well,  as  truly,  and 
as  faithfully  as  if  the  same  were  herein  particularly  described  and 
specifically  provided  for. 

The  engineer  shall  have  full  power  at  any  time  during  the  progress 
of  the  work  to  reject  any  materials  he  may  deem  unsuitable  for  the  pur- 
pose for  which  they  are  intended,  or  which  are  not  in  strict  conformity 
with  the  spirit  of  this  specification.  He  shall  also  have  the  power  to 
cause  any  inferior  or  unsafe  work  to  be  taken  down  and  altered  at  the 
cost  of  the  contractor. 

It  is  to  be  understood  that  the  final  inspection  and  acceptance  of 
the  work  are  to  take  place  at  the  building  after  erection,  and  that  any 
inspection  and  acceptance  of  material  and  workmanship  at  the  mills, 
shops,  etc.,  to  facilitate  the  progress  of  the  work,  shall  not  preclude 
rejection  at  the  building  if  the  same  be  unsuitable. 

*  Stevens  Indicator,  Vol.  18,  p.  373,  1901, 


TYPICAL  SPECIFICATIONS  761 

Any  disagreement  or  difference  between  the  purchaser  and  the  con- 
tractor, upon  any  matter  or  thing  arising  from  this  specification  or  the 
drawings  which  form  a  part  thereof,  shall  be  referred  to  the  engineer, 
whose  decision  and  interpretation  of  the  same  shall  be  considered  final, 
conclusive,  and  binding  upon  both  parties. 

Risk  and  Blame.  —  The  contractor  is  to  assume  all  risks  and  bear 
any  loss  occasioned  by  neglect  or  accident  during  the  progress  of  the 
work  until  the  same  shall  have  been  completed  and  accepted  by  the 
engineer. 

Permits.  —  The  contractor  must  pay  for  all  permits  and  inspec- 
tors' fees  or  any  other  charges  from  borough,  city,  county,  or  state 
officers. 

Dimensions.  —  During  the  progress  of  the  work  the  contractor  will 
be  required  to  keep  at  the  building  site  a  complete  set  of  drawings 
and  a  copy  of  this  specification.  The  contractor  will  consider  dimensions 
shown  on  drawings  to  be  approximate  as  to  location  of  machinery, 
but  sufficiently  accurate  for  the  purpose.  Absolute  dimensions  must 
be  gotten  from  the  location  of  the  boilers,  engines,  and  other  machinery 
after  they  are  set.  No  drawing  shall  be  scaled. 

The  purchaser  reserves  the  right  to  put  other  parties  at  work  on  the 
premises  erecting  machinery  and  doing  other  work  during  the  con- 
tinuance of  this  contract. 

The  contractor  must  conform  to  such  rules  and  regulations  as  are 
in  force  on  the  property  and  carry  out  this  contract  with  as  little 
interference  as  possible  with  the  other  work  of  the  purchaser. 

DRAWINGS.* 

The  following  is  a  list  of  the  drawings  which  accompany  this 
specification  and  which  form  a  part  thereof: 

(Title) No.  (Drawing  Number) 

do do etc. 

DESCRIPTION  OF  PLANT. 

Steam  Piping.  —  The  plant  consists  of  four  (4)  250-horse-power 
boilers  from  which  steam  will  be  led  to  two  (2)  independent  16-inch 
steam  headers.  Two  (2)  8-inch  take-offs  will  be  carried  from  each 
boiler,  one  (1)  to  each  header.  These  8-inch  take-offs  will  consist  of 
large  radius  wrought-iron  bends  of  a  quality  hereinafter  described. 
From  each  header  a  6-inch  supply  will  be  led  to  the  low-pressure 
cylinders  of  these  engines.  From  the  header  nearest  the  engine  room 
two  (2)  6-inch  supplies  will  be  led  through  the  partition  wall  into  the 
*  Drawings  have  been  omitted. 


762  STEAM  POWER  PLANT  ENGINEERING 

cellar  for  the  purpose  of  furnishing  steam  to  the  condenser  pumps. 
From  one  (1)  header  a  6-inch  and  5-inch  supply  will  be  carried  to  a 
tandem  compound  engine  and  thence  5-inch  to  the  two  (2)  exciter 
engines. 

The  condenser  steam  lines  will  be  continued  under  the  engine-room 
floor  so  as  to  form  a  reserve  steam  supply  to  the  exciter  engines. 
From  the  end  of  each  16-inch  main  steam  header  a  3-inch  loop  will  be 
carried  to  and  from  the  fire  and  boiler  feed  pumps.  The  general 
arrangement  with  sizes  of  pipes,  separators,  valves,  etc.,  is  shown  on 
drawing  No. . 

The  contractor's  high-pressure  steam  work  will  begin  with  the  steam 
nozzles  on  the  boilers  and  end  with  the  throttle  valves  of  pumps  and 
engines.  The  contractor  will  furnish  all  pump  throttle  valves  but  not 
engine  throttle  valves. 

Condenser  and  Exhaust  Piping.  —  The  three  (3)  jet  condensers  and 
air  pumps  will  be  set  under  engine-room  floor.  Injection  water  will 
be  led  to  the  condenser  through  a  16-inch  cast-iron  pipe  which  begins 
at  a  point  6  feet  outside  the  wall  of  the  boiler  house  and  runs  from 
this  point  to  the  condensers.  After  leaving  the  connections  for  the 
fourth  condenser,  this  main  reduces  to  14  inches  in  diameter;  after 
leaving  the  third  it  reduces  to  10  inches  in  diameter;  and  after 
leaving  the  second  it  reduces  to  3  inches  in  diameter. 

This  injection  water  main  will  be  connected  with  the  city  supply  by 
a  5-inch  cast-iron  water  pipe;  and  the  discharge  pipe  will  be  of  cast- 
iron  and  will  be  led  from  the  condensers,  beginning  with  a  diameter  of 
5  inches,  increasing  to  10  inches,  and  again  increasing  to  14  inches, 
and  then  to  16  inches  after  the  last  condenser,  from  whence  it  will  be 
carried  to  a  point  6  feet  outside  of  the  boiler-house  wall. 

The  low-pressure  cylinder  of  the  tandem  compound  engine  will  be 
connected  by  a  10-inch  and  a  12-inch  pipe  with  No.  1  condenser. 
Connection  will  be  made  between  both  high-  and  low-pressure  cylin- 
ders of  the  750-horse-power  engines  with  18-inch  condenser  exhaust 
headers.  Between  the  exhaust  header  and  the  cylinder  of  the  engines, 
connections  will  be  made  with  a  free  exhaust  main  leading  to  and 
from  exhaust  riser  extending  through  the  roof  of  the  boiler  house. 
The  general  arrangement  and  the  sizes  of  the  gate  valves,  free  exhaust 
valves,  etc.,  are  shown  on  drawing  No. . 

The  exhaust  from  the  two  (2)  exciter  engines  and  the  three  (3) 
condensers,  the  boiler  feed  pumps,  and  the  fire  pumps  will  be  collected 
and  led  to  a  1500-horse-power  open-type  feed- water  heater  located  in 
the  boiler  room.  This  heater  will  be  provided  with  an  exhaust  riser- 
and  exhaust  head  extending  through  the  roof  of  the  boiler  house. 


TYPICAL  SPECIFICATIONS  763 

The  arrangement  of  the  auxiliary  exhaust  pipe  is  shown  on  drawing 
No. . 

Boiler  Feed  Water  Piping.  —  To  the  feed-water  heater  above 
mentioned  a  2J-inch  supply  of  fresh  water  will  be  furnished.  The 
boiler  feed  pumps  will  draw  the  feed  water  from  the  heater  through  a 
6-inch  cast-iron  suction  pipe.  They  will  deliver  it  through  3-inch 
brass  pipes  to  two  (2)  hot-water  meters  equipped  with  by-passes  and 
thence  through  a  3-inch  brass  pipe  line  to  boilers.  A  reserve  wrought- 
iron  3-inch  feed  pipe  line  will  be  run  parallel  to  the  brass  feed  pipe 
line  and  so  connected  up  to  it  that  it  may  be  used  as  a  reserve  in  case 
of  accident  to  the  brass  line.  Connections  from  3-inch  lines  to  boilers 
will  be  2J-inch  brass  pipe. 

The  general  arrangement  of  the  feed  pipe,  including  gate  valves, 
check  valves,  sizes,  etc.,  is  shown  on  drawing  No. . 

Drip  Piping. — There  will  be  two  (2)  systems  of  drip  piping, 
including  traps,  trap  by-passes,  valves,  etc. 

The  high-pressure  system  comprises  all  the  drips  from  the  main 
steam  header,  separators,  and  high-pressure  steam  lines.  The  water 
from  these  traps  will  be  led  to  the  feed-water  heater,  drawing 
No. . 

The  low-pressure  system  comprises  all  the  drips  from  exhaust  mains, 
receivers,  and  cylinder  drain  cocks. 

This  water  will  be  led  to  a  catch-basin  from  whence  it  will  be 
pumped  by  a  tank  pump.  This  pump  will  be  furnished  by  the  pur- 
chaser, drawing  No. . 

Blow-off  Pipes.  —  A  separate  and  distinct  wrought-iron,  2J-inch 
blow-off  pipe  will  be  run  from  the  blow-off  valve  of  each  of  the 
four  (4)  water-tube  boilers.  These  pipes  will  be  run  in  trenches  in 
the  boiler-room  floor  to  catch-basins  located  outside  of  building.  The 
purchaser  will  construct  the  trenches  and  catch-basins. 

Water  Piping.  —  The  purchaser  will  make  a  6-inch  connection  from 
the  city  main  to  the  northwest  corner  of  the  boiler  room.  From  this 
point  the  contractor  will  run  6-inch  cast-iron  pipe  to  the  inlet  of 
the  fire  pump.  From  this  he  will  run  6-inch  pipe  to  the  several  fire 
hydrants.  He  will  connect  this  6-inch  line  with  the  injection  water 
main  with  a  5-inch  cast-iron  connection.  Near  the  fire  pump  he  will 
run  a  2J-inch  connection  to  the  feed-water  heater.  The  pipe  around 
the  fire  pump  will  be  so  arranged  that  ordinarily  city  pressure  will  be 
maintained  on  the  water  system,  but  in  case  of  need  the  fire  pump  may 
be  used  to  increase  this,  pressure. 
•  This  piping  is  all  shown  on  drawing  No. . 

Wrought-iron  Pipe.  —  All  wrought-iron  pipes  or  steam  lines  to  be 


764 


STEAM   POWER   PLANT   ENGINEERING 


guaranteed  full-weight  lap-welded  wrought-iron  pipe,  in  accordance 
with  the  standard  dimensions  as  given  in  the  table  below.  But  the 
16-inch  headers  will  be  16  inches  O.D.,  15.04  inches  I.D.,  and  shall 
weigh  87  pounds  per  foot. 


Nominal  Internal. 

Actual  External. 

Actual  Internal. 

Nominal  Weight  per 
Foot. 

Inches. 

Inches. 

Inches. 

Pounds. 

1 

1.315 

1.048 

1.663 

H 

1.66 

1.38 

2.444 

if 

1.9 

1.611 

2.678 

2 

2.375 

2.067 

3.609 

% 

2.875 

2.468 

5.739 

3 

3.5 

3.067 

7.536 

3* 

4 

3.548 

9.001 

4 

4.5 

4.026 

10.665 

4* 

5 

4.508 

12.34 

5 

5.563 

5.045 

14.502 

6 

6.624 

6.065 

18.762 

8 

8.625 

7.982 

28.35 

10 

10.75 

10.019 

40.065 

14 

15 

14.25 

57.893 

15 

16 

15.25 

71.77 

All  pipes  carrying  hot  water,  such  as  feed-water,  blow-offs,  and 
high-pressure  drip  pipes,  are  to  be  extra  heavy. 

All  lap-welded  pipe  shall  be  proved  to  500  pounds  pressure  per 
square  inch  before  shipment.  Butt- welded  pipe  shall  be  proved  to 
300  pounds  pressure  per  square  inch  before  shipment. 

Exhaust  pipes  are  to  be  wrought  iron  or  steel  in  every  case;  sizes 
8  inches  and  above  may  be  light  O.D.  tubing  with  flanges  peened 
or  expanded  on,  but  must  be  absolutely  tight  at  28  inches  vacuum. 
These  pipes  must  be  tested  to  25  pounds  hydrostatic  pressure  after 
erection. 

All  high-pressure  wrought-iron  pipe  when  in  place  is  to  be  tested  at 
250  pounds  air  pressure,  and  must  be  good  for  a  working  pressure  of 
175  pounds.  All  high-pressure  steam  pipe  to  be  tested  after  erection 
at  175  pounds  steam  pressure,  to  the  satisfaction  of  the  engineer  in 
every  particular. 

ts/Cast-Iron  Pipe.  —  The  cast-iron  pipe  for  injection  and  discharge 
water  and  for  all  other  purposes,  as  shown  on  the  accompanying 
drawings,  shall  be  made  of  tough  gray  iron  not  less  than  f  inch  thick 
at  any  point,  free  from  blowholes,  true  to  pattern,  and  of  workman- 
like finish.  It  shall  be  tested  at  250  pounds  pressure  before  erection, 
and  when  in  place  shall  be  tested  to  25  pounds  hydrostatic  pressure. 

Cast-iron  pipes  inside  of  buildings  are  to  be  flanged  with  flanged 
fittings. 


TYPICAL  SPECIFICATIONS 


765 


Outside  of  buildings  they  shall  be  bell  and  spigot  ends,  and  must  be 
coated  with  tar  both  inside  and  outside. 

j^gJjjgSj — •  All  steam  valves  over  two  inches,  except  throttle  valves, 
will  be  gate  valves.  All  gate  valves  will  be  Chapman's  best  make  or 
equivalent. 

On  high-pressure  steam  lines,  valves  must  be  fitted  with  bronze  re- 
movable seats,  outside  screw  and  yoke,  and  by-passed  from  5  inches  up. 

On  exhaust  lines,  standard  pattern  gate  valves  will  be  used  (Chap- 
man or  equivalent),  with  inside  screw. 

Globe  valves  on  high-pressure  steam  and  water  lines  will  be  Schutte's 
make  or  equivalent. 

Free  exhaust  valves  will  be  either  Schutte's  make  or  equivalent. 

Exhaust  valves  between  engines  and  condensers,  injection  water 
valves  at  condensers,  as  well  as  priming  water  valves,  will  have  their 
stems  extended  and  fitted  with  floor  stands,  so  that  they  can  be 
operated  from  floor  above.  Floor  stands  will  be  polished  all  over, 
with  polished  hand  wheels  and  indicators. 

All  high-pressure  valves  will  be  tested  to  the  satisfaction  of  the 
engineer,  at  a  hydrostatic  pressure  of  350  pounds  per  square  inch 
and  air  pressure  of  100  pounds  between  gates. 

Standard  valves  for  exhaust  and  water  will  be  tested  to  150  pounds. 

Flanges  and  Fittings.  —  All  flanges  and  fittings  for  high-pressure 
steam  lines  3  inches  and  over  will  be  open-hearth  steel  castings. 
These  castings  must  be  perfectly  solid,  made  from  heavy  patterns,  and 
free  from  blowholes  or  other  defects.  They  must  be  tested  at  works 
to  500  pounds  hydrostatic  pressure,  and  guaranteed  for  a  working 
steam  pressure  of  175  pounds  per  square  inch. 

The  diameter,  thickness,  and  drilling  of  flanges  must  not  be  less  than 
the  dimensions  given  in  the  following  table. 


Size  of  Pipe. 

Diameter  of 
Flange. 

Thickness. 

Number  of 
Bolts. 

Size  of  Bolts. 

Inches. 

Inches. 

Inches. 

Inch. 

16 

25 

If 

20 

1 

14 

23 

If 

16 

1 

12 

20 

H 

16 

1 

10 

17} 

if 

12 

1 

9 

16 

12 

1 

8 

15 

if 

12 

7 

14 

H 

12 

1 

6 

13 

8 

5 

11 

H 

8 

.4} 

10} 

i& 

8 

4 

10 

8 

3} 

9 

i 

8 

3 

9 

1 

8 

766  STEAM  POWER  PLANT  ENGINEERING 

Below  3  inches,  on  all  high-pressure  steam,  water,  or  drip  lines,  extra 
heavy  screwed  cast-iron  fittings  are  to  be  used,  with  sufficient  extra 
heavy  flange  unions,  so  that  any  section  of  pipe  can  be  readily  taken 
out  without  disturbing  the  balance. 

The  brass  boiler  feed  piping  will  be  made  up  with  extra  heavy  brass 
screwed  fittings  made  from  cast-iron  patterns. 

Pipe  will  be  iron  pipe  size. 

Brass  flange  unions  will  be  made  from  standard  cast-iron  patterns, 
and  a  sufficient  number  used  to  make  up  the  sections  readily. 
-  On  low-pressure  piping,  standard  fittings  and  flanges  are  to  be  used. 
Sizes  6  inches  and  above  are  to  be  flanged,  and  below  6  inches,  screwed, 
with  sufficient  number  of  flange  unions  in  same. 

All  high-pressure  flanges  are  to  be  recessed  on  face,  pipe  is  to  be 
screwed,  then  peened  up  tight,  and  must  stand  the  tests  given 
above. 

Standard  flanges  must  stand  the  tests  as  stated. 

All  bolt  holes  must  be  drilled  in  solid  metal;  no  cored  holes  will  be 
allowed. 

On  blow-off  lines,  long  sweep  fittings  are  to  be  used,  with  extra 
heavy  flange  unions. 

Gaskets  and  Bolts.  —  All  gaskets  on  high-pressure  steam  lines  must 
be  copper. 

On  exhaust  and  water  lines  metallic  gaskets  are  to  be  used. 

All  bolts  must  have  hexagonal  heads  and  nuts,  no  matter  what  size, 
points  to  be  finished  and  nuts  to  be  cold  punched. 

Pipe  Covering.  —  All  high-pressure  steam  lines,  separators,  and  all 
drip  lines  between  high-pressure  steam  lines  and  their  traps  will  be 
covered  with  an  approved  magnesia  pipe  covering.  All  fittings  will 
have  molded  magnesia  coverings  to  fit  them  neatly. 

Exhaust  Heads.  —  The  contractor  will  furnish  and  erect  two  (2) 
exhaust  heads,  one  on  the  free  exhaust  from  the  engines  and  the  other 
on  the  free  exhaust  from  the  feed- water  heater. 

Pipe  Supports.  —  There  will  be  furnished  and  erected  the  necessary 
wrought-iron  hangers  to  support  the  two  (2)  16-inch  headers  in  the 
boiler  room.  There  will  also  be  supplied  the  necessary  supports  for 
the  steam  mains  to  the  pumps  and  exciter  engines.  The  injection 
water  main  and  the  condenser  exhaust  header  will  be  hung  from  the 
engine  room  floor  beams.  The  free  exhaust  header  and  the  discharge 
water  main  will  be  supported  on  piers  and  saddles  from  the  cellar 
floor.  The  saddles  to  be  furnished  by  the  contractor. 

Separators.  —  The  contractor  will  furnish  and  erect,  as  shown  on 
drawing  No. ,  eleven  (11)  steam  separators. 


TYPICAL  SPECIFICATIONS 


767 


Anchors.  —  All  live-steam  mains,  and  especially  the  16-inch  steam 
headers  in  the  boiler  room,  will  be  anchored  so  as  to  secure  no  move- 
ment (at  point  of  anchorage)  on  account  of  the  expansion  and  con- 
traction or  vibration.  Anchors  to  be  placed  where  needed  after  the 
rest  of  the  work  is  completed. 

Relief  Valves.  —  The  boiler-feed  and  fire  pumps  shall  be  fitted 
with  automatic  controlling  devices  so  that  a  certain  definite  maxi- 
mum pressure  may  be  maintained  on  the  feed- water  and  fire 
systems. 

Overflow  Pipes.  —  The  contractor  will  furnish  an  overflow  relief  valve 
for  the  fire  pump  and  will  connect  same  with  the  drain  outside  the  boiler 
house.  He  will  also  connect  the  blow-off  and  overflow  from  the  feed- 
water  heater  with  the  said  drain. 

Long  Radius  Bends.  —  Where  shown  on  the  drawings,  changes  in 
direction  of  pipe  runs  will  be  made  with  long  radius  bends.  .  These 
bends  will  be  of  wrought-iron  of  the  quality  described  above.  Nine 
inches  of  straight  pipe  will  be  left  on  the  ends  of  the  bends  to  cut 
necessary  threads. 

The  following  table  shows  the  minimum  radii  for  these  wrought-iron 
bends. 


Size. 

Minimum  Radius. 

Size. 

Maximum 

Radius. 

Inches. 

Inches. 

Inches. 

Feet. 

Inches. 

2 

4* 

7 

3 

0 

2* 

6 

8 

3 

4 

3 

8 

9 

4 

0 

3| 

10 

10 

4 

4 

4 

14 

12 

5 

6 

4* 

16 

14 

7 

0 

5 

20 

16 

7 

6 

6 

24 

I -Beam  Supports.  —  The  contractor  will  furnish  and  erect  on  the 
top  chord  of  the  roof  trusses  in  the  boiler  room  a  sufficient  length  of 
I  beams  or  channels  from  which  he  will  hang  the  16-inch  headers. 
The  steam  connections  between  16-inch  header  and  the  engines  will  be 
supported  by  underneath  rod  trusses,  and  vibration  will  be  taken  up 
by  means  of  lateral  ties. 

Flange  Covering.  —  After  all  the  other  work  is  completed  and  the 
plant  has  been  in  operation  not  less  than  two  (2)  weeks,  the  con- 
tractor will  cover  all  flanges  and  other  joints  with  an  approved 
magnesia  covering. 


768  STEAM  POWER  PLANT  ENGINEERING 

Reducing  and  Relief  Valves.  —  The  contractor  will  furnish  and  erect 
two  (2)  reducing  valves  on  the  throttle  valves  of  the  low-pressure 
cylinders  of  the  750-horse-power  engines.  These  valves  will  be  of 
the  Kiely  type  or  equivalent. 

There  will  also  be  furnished  and  erected  the  necessary  traps  from 
the  receivers  between  the  high  and  low-pressure  cylinders. 

Check  Valves.  —  All  drip  lines  will  be  equipped  with  check  valves 
which  will  be  set  at  the  lowest  point  in  the  line.  Two  (2)  three-inch 
brass  checks  will  be  placed  on  the  boiler  feed  pump  outlets,  all  as 
shown  on  drawing  No. . 

The  outlet  of  the  boiler  feed  pumps  will  also  have  a  relief  valve 
which  will  return  the  water  to  the  feed-water  heater  in  case  the  pres- 
sure on  the  pump  outlet  should  rise  above  a  certain  definite  point. 

Traps.  —  All  traps  on  high-pressure  drip  lines  will  be  extra  heavy 
steam  traps. 

Painting.  —  All  pipes  shall  be  painted  with  two  (2)  coats  of  slate 
graphite  and  linseed  oil,  or  other  good  pipe  paint  satisfactory  to  the 
engineer. 

The  exhaust  and  condenser  piping  must  be  thoroughly  painted  twice 
under  vacuum. 

Damper  Regulator.  —  The  purchaser  will  furnish  the  damper  regu- 
lator, but  the  contractor  will  set  it  and  connect  it  with  both  steam 
headers,  water  supply,  and  damper. 

Shields.  —  Where  steam  mains  pass  through  the  partition  wall  of 
the  building  the  contractor  will  neatly  close  the  opening  with  a  sheet- 
metal  shield  so  as  to  prevent  dust  from  the  boiler  room  from  entering 
the  engine  room. 

Meters.  —  The  contractor  will  furnish  and  connect  as  shown  on 

drawing  No. two  (2)  3-inch  hot-water  meters,  approved  by  the 

engineer,  to  have  a  capacity  of  200  gallons  per  minute. 

Gauges.  —  The  purchaser  will  furnish  the  customary  gauges  and 
boards,  which  will  be  set  up  and  connected  by  the  contractor. 

Machinery. — The  purchaser  will  furnish  all  engines,  pumps,  con- 
densers, feed-water  heater,  boilers,  foundations,  but  the  contractor 
shall  connect  up  the  above  machinery  according  to  the  evident  intent 
and  meaning  of  this  specification. 

The  purchaser  will  not  furnish  any  pipe,  valves,  fittings,  pipe  cover- 
ing, etc.,  or  anything  connected  with  the  work  except  the  machinery 
mentioned  above. 


TYPICAL  SPECIFICATIONS  769 

416.    Government  Specification  and  Proposal  for  Supplying  Coal. 

U.  S.  TREASURY  DEPARTMENT. 

United  States 

,  190.. 

PROPOSAL. 

1  Sealed  proposals  will  be  received  at  this  office  until  2  o'clock  p.  m., 

2     ,  190 . . ,  for  supplying  coal  to  the  United  States 

3 building  at 

4     as  follows : 

5     

6     

7     

8  The  quantity  of  coal  stated  above  is  based  upon  the  previous  annual 

9  consumption,  and  proposals  must  be  made  upon  the  basis  of  a  delivery  of 

10  10  per  cent  more  or  less  than  this  amount,  subject  to  the  actual  requirements 

11  of  the  service 

12  Proposals  must  be  made  on  this  form,  and  include  all  expenses  incident 

13  to  the  delivery  and  stowage  of  the  coal,  which  must  be  delivered  in  such 

14  quantities,  and  at  such  times  within  the  fiscal  year  ending  June  30,  190  , 

15  as  may  be  required. 

16  Proposals  must  be  accompanied  by  a  deposit   (certified  check,  when 

17  practicable,  in  favor  of ) 

18  amounting  to  10  per  cent  of  the  aggregate  amount  of  the  bid  submitted,  as 

19  a  guaranty  that  it  is  bona  fide.     Deposits  will  be  returned  to  unsuccessful 

20  bidders  immediately  after  award  has  been  made,  but  the  deposit  of  the 

21  successful  bidder  will  be  retained  until  after  the  coal  shall  have  been 

22  delivered,  and  final  settlement  made  therefor,  as  security  for  the  faithful 

23  performance  of  the  terms  of  the  contract,  with  the  understanding  that  the 

24  whole  or  a  part  thereof  may  be  used  to  liquidate  the  value  of  any  deficiencies 

25  in  quality  or  delivery  that  may  arise  under  the  terms  of  the  contract. 

26  When  the  amount  of  the  contract  exceeds  $10,000,  a  bond  may  be  exe- 

27  cuted  in  the  sum  of  25  per  cent  of  the  contract  amount,  and  in  this  case,  the 

28  deposit  or  certified  check  submitted  with  the  proposal  will  be  returned  after 

29  approval  of  the  bond. 

30  The  bids  will  be  opened  in  the  presence  of  the  bidders,  their  representa- 

31  tives,  or  such  of  them  as  may  attend,  at  the  time  and  place  above  specified. 

32  In  determining  the  award  of  the  contract,  consideration  will  be  given  to 

33  the  quality  of  the  coal  offered  by  the  bidder,  as  well  as  the  price  per  ton, 

34  and  should  it  appear  to  be  to  the  best  interests  of  the  Government  to 

35  award  the  contract  for  supplying  coal  at  a  price  higher  than  that  named  in 

36  lower  bid  or  bids  received,  the  award  will  be  so  made. 

37  The  right  to  reject  any  or  all  bids  and  to  waive  defects  is  expressly 

38  reserved  by  the  Government. 


770  STEAM   POWER  PLANT  ENGINEERING 


DESCRIPTION  OF  COAL  DESIRED.* 

39  Bids  are  desired  on  coal  described  as  follows : 

40  

41  

42 

43  

44 

45  

46 

47  

48  

49  

50  Coals  containing  more  than  the  following  percentages,  based  upon  dry 

51  coal,  will  not  be  considered: 

52  Ash per  cent. 

53  Volatile  matter per  cent. 

54  Sulphur per  cent. 

55  f  Dust  and  fine  coal  as  delivered  at  point  of  consumption per  cent. 

DELIVERY. 

56  The  coal  shall  be  delivered  in  such  quantities  and  at  such  times  as  the 

57  Government  may  direct. 

58  In  this  connection,  it  may  be  stated  that  all  the  available  storage  capacity 

59  of  the  coal  bunkers  will  be  placed  at  the  disposal  of  the  contractor  to 

60  facilitate  delivery  of  coal  under  favorable  conditions. 

61  After  verbal  or  written  notice  has  been  given  to  deliver  coal  under  this 

62  contract,  a  further  notice  may  be  served  in  writing  upon  the  contractor  to 

63  make  delivery  of  the  coal  so  ordered  within  twenty-four  hours  after  receipt 

64  of  said  second  notice. 

65  Should  the  contractor,  for  any  reason,  fail  to  comply  with  the  second 

66  request  the  Government  will  be  at  liberty  to  buy  coal  in  the  open  market, 

67  and  to  charge  against  the  contractor  any  excess  in  price  of  coal  so  purchased 

68  over  the  contract  price. 

SAMPLING. 

69  Samples  of  the  coal  delivered  will  be  taken  by  a  representative  of  the 

70  Government. 

71  In  all  cases  where  it  is  practicable,  the  coal  will  be  sampled  at  the  time 

*  NOTE. — This  information  will  be  given  by  the  Government  as  may  be  deter- 
mined by  boiler  and  furnace  equipment,  operating  conditions,  and  the  local  market, 

t  NOTE.  —  All  coal  which  wilJ  pass  through  a  £-inch  round-hole  screen. 


TYPICAL  SPECIFICATIONS       .  771 

72  it  is  being  delivered  to  the  building.     In  case  of  small  deliveries,  it  may  be 

73  necessary  to  take  these  samples  from  the  yards  or  bins.     The  sample 

74  taken  will  in  no  case  be  less  than  the  total  of  one  hundred  (100)  pounds,  to 

75  be  selected  proportionally  from  the  lumps  and  fine  coal,  in  order  that  it 

76  will  in  every  respect  truly  represent  the  quality  of  coal  under  considera- 

77  tion. 

78  In  order  to  minimize  the  loss  in  the  original  moisture  content  the  gross 

79  sample  will  be  pulverized  as  rapidly  as  possible  until  none  of  the  fragments 

80  exceed  J  inch  in  diameter.     The  fine  coal  will  then  be  mixed  thoroughly 

81  and  divided  into  four  equal  parts.     Opposite  quarters  will  be  thrown  out, 

82  and  the  remaining  portions  thoroughly  mixed  and  again  quartered,  throw- 

83  ing  out  opposite  quarters  as.  before.     This  process  will  be  continued  as 

84  rapidly  as  possible  until  the  final  sample  is  reduced  to  such  amount  that  all 

85  of  the  final  sample  thus  obtained  will  be  contained  in  the  shipping  can  or 

86  jar  and  sealed  air-tight. 

87  The  sample  will  then  be  forwarded  to  the  Chief  Clerk  of  the  Treasury 

88  Department,  care  of  the  storekeeper. 

89  If  desired  by  the  coal  contractor,  permission  will  be  given  to  him,  or  his 

90  representative,  to  be  present  and  witness  the  "quartering  and  preparation  of 

91  the  final  sample  to  be  forwarded  to  the  Government  laboratories. 

92  Immediately  on  receipt  of  the  sample,  it  will  be  analyzed  and  tested  by 

93  the  Government,  following  the  method  adopted  by  the  American  Chemical 

94  Society,  and  using  a  bomb  calorimeter.     A  copy  of  the  result  will  be  mailed 

95  to  the  contractor  upon  the  completion  thereof. 


CAUSES   FOR    REJECTION. 

96  A  contract  entered  into  under  the  terms  of  this  specification  shall  not 

97  be  binding  if,  as  the  result  of  a  practical  service  test  of  reasonable  duration, 

98  the  coal  fails  to  give  satisfactory  results  due  to  excessive  clinkering,  or  to 

99  a  prohibitive  amount  of  smoke. 

100  It  is  understood  that  the  coal  delivered  during  the  year  will  be  of  the 

101  same  character  as  that  specified  by  the  contractor.     It  should,  therefore, 

102  be  supplied,  as  nearly  as  possible,  from  the  same  mine  or  group  of  mines. 

103  Coal  containing  percentages  of  volatile  matter,  sulphur,  and  dust  higher 

104  than  the  limits  indicated  on  line  54,  and  coal  containing  a  percentage  of 

105  ash  in  excess  of  the  maximum  limits  indicated  in  the  following  table  will 

106  be  subject  to  rejection. 

107  In  the  case  of  coal  which  has  been  delivered  and  used  for  trial,  or  which 

108  has  been  consumed  or  remains  on  the  premises  at  the  time  of  the  deter- 

109  mination  of  its  quality,  payment  will  be  made  therefor  at  a  reduced  price 

110  computed  under  the  terms  of  this  specification. 

111  Occasional  deliveries  containing  ash  up  to  the  percentage  indicated  in 

112  the  column  of  "Maximum  limits  for  ash,"  on  page  700,  may  be  accepted. 


772 


STEAM  POWER  PLANT  ENGINEERING 


113  Frequent  or  continued  failure  to  maintain  the  standard  established  by 

114  the  contractor,  however,  will  be  considered  sufficient  cause  for  cancellation 

115  of  the  contract. 

*  PRICE  AND  PAYMENT. 

116  Payment  will  be  made  on  the  basis  of  the  price  named  in  the  proposal 

117  for  the  coal  specified  therein,  corrected  for  variations  in  heating  value  and 

118  ash,  as  shown  by  analysis,  above  and  below  the  standard  established  by 

119  contractor  in  this  proposal.     For  example,  if.  the  coal  contains  two  (2) 

120  per  cent,  more  or  less,  British  thermal  units  than  the  established  standard, 

121  the  price  will  be  increased  or  decreased  two  (2)  per  cent  accordingly. 

122  The  price  will  also  be  further  corrected  for  the  percentages  of  ash.     For 

123  all  coal  which  by  analysis  contains  less  ash  than  that  established  in  this 

124  proposal  a  premium  of  1  cent  per  ton  for  each  whole  per  cent  less  ash  will 

125  be  paid.     An  increase  in  the  ash  content  of  two  (2)  per  cent  over  the 

126  standard  established  by  contractor  will  be  tolerated  without  exacting  a 

127  penalty  for  the  excess  of  ash.     When  such  excess  exceeds  two  (2)  per  cent 

128  above  the  standard  established,  deductions  will  be  made  from  price  paid 

129  per  ton  in  accordance  with  following  table : 


Ash  as  estab- 
lished in 
proposal. 

No 

deduc- 
tion for 
limits 
below. 

Cents  per  ton  to  be  deducted. 

Maxi- 
mum 

limits 
for 
ash. 

12 
13 
14 
14 
15 
16 
16 
17 
18 
19 
19 
20 
21 
22 

2 

4 

7 

12 

18 

25 

35 

Percentages  of  ash  in  dry  coal. 

Per  cent. 
5  
6        

7 
8 
9 
10 
11 
12 
13 
14 
15 
16 
17 
18 
19 
20 

7-  8 
8-  9 
9-10 
10-11 
11-12 
12-13 
13-14 
14-15 
15-16 
16-17 
17-18 
18-19 
19-20 
20-21 

8-  9 
9-10 
10-11 
11-12 
12-13 
13-14 
14-15 
15-16 
16-17 
17-18 
18-19 
19-20 
20-21 
21-22 

9-10 
10-11 
11-12 
12-13 
13-14 
14-15 
15-16 
16-17 
17-18 
18-19 

10-11 
11-12 
12-13 
13-14 
14-15 
15-16 

11-12 
12-13 
13-14 

12-13 
13-14 
14-15 
15-16 
16-17 
17-18 
18-19 
19-20 
20-21 
21-22 

13-14 
14-15 
15-16 
16-17 
17-18 

7  

g 

14-15 
15-16 
16-17 

9  

10  

11 

16-17 
17-18 
18-19 
19-20 

17-18 
18-19 
19-20 
20-21 
21-22 
22-23 



12 

13   

14  

15 

19-20 
20-21 
21-22 
22-23 

20-21 
21-22 
22-23 

16  
17 



....... 

18 

*  NOTE.  —  The  economic  value  of  a  fuel  is  affected  by  the  actual  amount  of  com- 
bustible matter  it  contains,  as  determined  by  its  heating  value  shown  in  British 
thermal  units  per  pound  of  fuel,  and  also  by  other  factors,  among  which  is  its  ash 
content.  The  ash  content  not  only  lowers  the  heating  value  and  decreases  the 
capacity  of  the  furnace,  but  also  materially  increases  the  cost  of  handling  the  coal, 
the  labor  of  firing,  and  the  cost  of  the  removal  of  ashes,  etc. 


TYPICAL  SPECIFICATIONS  773 

Proposals  to  receive  consideration  must  be  submitted  upon  thin  form  and  contain 
all  of  the  information  requested. 


,  190 

The  undersigned  hereby  agree  to  furnish  to  the  U.  S 

building  at ,  the  coal  described,  in  tons 

of  2,240  pounds  each  and  in  quantity,  10  per  cent  more  or  less  than  that  stated 
on  page  697,  as  may  be  required  during  the  fiscal  year  ending  June  30,  190  , 
in  strict  accordance  with  this  specification;  the  coal  to  be  delivered  in  such 
quantities  and  at  such  times  as  the  Government  may  direct. 

Price  per  ton  (2,240  pounds) $ 

Commercial  name  of  the  coal    

Name  of  the  mine  or  mines 

Location  of  the  mine  or  mines 

Name  or  other  designation  of  the  coal  bed  or  vein 

Size  (indicate  information  which  will  apply)  — 

Unsized Lump Run  of  mine 

( Round )~ 

Screened,    through inch    and    over inch  i  Square  ) 

iBar  screen. 
Data  to  establish  a  basis  for  payment: 

British  thermal  units  in  coal  as  delivered 

Ash  in  dry  coal  (Method  of  American  Chemical  Society) per  cent. 

It  is  important  that  the  above  information  does  not  establish  a  higher  standard  than  can  be 
actually  maintained  under  the  terms  of  the  contact;  and  in  this  connection  it  should  be  noted 
that  the  small  samples  taken  from  the  mine  are  invariably  of  higher  quality  than  the  coal  actually 
delivered  therefrom.  It  is  evident,  therefore,  that  it  will  be  to  the  best  interests  of  the  contractor 
to  furnish  a  correct  description  with  average  values  of  the  coal  offered,  as  a  failure  to  maintain  the 
standard  established  by  contractor  will  result  in  deductions  from  the  contract  price,  and  may 
cause  a  cancellation  of  the  contract,  while  deliveries  of  a  coal  of  higher  grade  than  quoted  will  be 
paid  for  at  an  increased  price. 


Signature : . . . . 
Address ; 


Name  of  corporation,    

Name  of  president,   

Name  of  secretary,   

Under  what  law  (State)  corporation  is  organized 


CHAPTER  XX. 

A  TYPICAL  CENTRAL  STATION. 
Steam  Turbines.  —  Alternating  Currents. 

The  Fisk  Street  Station  of  the  Commonwealth  Edison  Company, 
Chicago,  is  an  excellent  example  of  modern  central  station  practice. 
The  present  (June,  1910)  rated  capacity  of  the  plant  is  120,000  kilo- 
watts, though  space  is  available  for  a  considerable  increase.  The 
station  is  located  on  the  banks  of  the  Chicago  River  near  Fisk  and 
Twenty-second  streets,  as  indicated  in  Fig.  438,  and  is  about  one  and 
one-half  miles  south  of  the  center  of  distribution  of  the  present  load. 
The  location  of  the  station  between  the  east  and  west  slips  of  the  river 
secures  an  unusual  advantage  in  the  location  of  the  intake  and  discharge 
tunnels,  and  the  extent  of  property  affords  ample  storage'  capacity  for 
coal.  Both  the  Chicago  &  Alton  and  the  Chicago,  Burlington  &  Quincy 
railways  extend  into  the  property,  giving  excellent  facilities  for  the 
transportation  of  coal,  ashes,  construction  materials,  and  machinery. 
The  plant  is  constructed  on  the  unit  basis,  each  turbine  and  generator 
having  its  own  boilers,  auxiliaries,  and  piping  system,  thus  permitting 
any  unit  to  be  shut  down  without  interfering  with  the  operation  of  the 
rest  of  the  system. 

Building.  —  The  main  building  rests  on  piles,  driven  to  hard  pan, 
capped  with  a  grillage  of  I  beams  and  concrete.  The  walls  are  of  red 
pressed  brick  trimmed  with  white  Bedford  stone.  The  windows  are 
25  feet  wide  and  32  feet  high,  the  sections  of  which  are  operated  by 
compressed  air.  Fig.  439  gives  a  view  of  the  north  elevation.  Large 
skylights  afford  ample  light  and  ventilation.  The  entire  interior  wall 
surface  of  the  turbine  room  is  finished  with  white  enameled  brick 
trimmed  with  terra  cotta.  The  boiler-room  walls  have  an  eight-foot 
wainscoting  of  enameled  brick,  the  remainder  being  red  pressed  brick. 
The  floors  are  of  concrete,  that  in  the  turbine  room  being  covered  with 
two-inch  hexagonal  terra-cotta  tile.  The  roofs  are  of  Roebling  concrete. 
The  total  width  of  the  building  is  243  feet,  the  turbine  room  taking 
up  61  feet,  the  boiler  room  142  feet,  and  a  car  track  the  remainder. 
A  50-ton  motor-driven  crane  spans  the  turbine  room  and  is  used  in 
connection  with  units  1  to  4,  and  a  60-ton  crane  is  provided  for  the 

774 


A  TYPICAL  CENTRAL  STATION 


775 


776 


STEAM  POWER  PLANT  ENGINEERING 


remaining  units.     A  5-ton  auxiliary  hoist  is  also  provided  on  the  main 

cranes.  In  the  boiler  room  a 
small  hand-power  crane  serves 
each  two  batteries  of  boilers. 

Coal  and  Ash  Handling. — 
An  interior  shed  extends  the 
entire  length  of  the  east  end 
of  the  building,  as  indicated 
in  Figs.  440  and  441.  Coal 
is  brought  in  on  cars  and 
dumped  or  shoveled  into  a 
track  hopper,  from  which  it 
is  delivered  to  the  overhead 
bunkers  by  the  conveying 
system.  A  crusher  is  placed 
$  between  the  track  hopper  and 
main  conveyor  to  be  used  in 
2  case  lump  coal  is  furnished. 
«  These  bunkers  have  a  capacity 
'•§  of  1200  tons  each,  sufficient 
.2  for  several  days  run.  The 
^  conveyors  are  driven  by  a 
o  15-horse-power  motor  and  are 
of  the  McCaslin  pattern,  end- 
§  less  chain,  with  overlapping 
g  buckets,  each  bucket  having 
a  capacity  of  100  pounds. 
The  conveyors  move  at  a  rate 
of  50  feet  per  minute,  giving 
a  service  capacity  of  75  tons 
per  hour  for  each  unit.  A 
separate  conveyor  and  bunker 
is-  installed  in  each  section  of 
16  boilers.  The  coal  bunkers 
feed  through  flexible  down 
spouts  to  the  stoker  maga- 
zines. Underneath  the  front 
end  of  the  stoker  is  a  fine-coal 
hopper  which  collects  the  fine 
coal  falling  through  the  grate 
and  discharges  it  into  the  conveyor  system,  as  in  Fig.  99.  The  ashes 
collect  in  the  ash  pit,  from  which  they  are  dumped  into  the  conveyor 


A  TYPICAL  CENTRAL  STATION 


777 


778 


STEAM  POWER  PLANT  ENGINEERING 


A  TYPICAL  CENTRAL  STATION  779 

and  carried  to  an  ash  bin  directly  over  the  coal  track.  Illinois  screen- 
ings furnish  the  greater  part  of  the  fuel.  Provision  is  also  made  for 
outside  storage. 

Boilers.  —  The  boiler  plant  is  divided  into  five  sections,  each  section 
consisting  of  sixteen  500-horse-power  B.  &  W.  boilers  arranged  in  bat- 
teries of  eight  and  equipped  with  B.  &  W.  chain  grates.  The  settings 
are  installed  back  to  back,  as  illustrated  in  Fig.  440.  Each  boiler  has 
two  42-inch  steam  drums,  approximately  5000  square  feet  of  heating 
surface,  and  about  1000  square  feet  of  superheating  surface.  Steam 
is  generated  at  a  pressure  of  200  pounds  per  square  inch  with  super- 
heat of  150  degrees  F.  The  ratio  of  water-heating  surface  to  grate 
surface  is  approximately  55  to  1,  and  the  ratio  of  the  total  heating 
surface  to  grate  surface  is  about  66  to  1.  When  burning  Illinois  screen- 
ings, an  average  depth  of  7  inches  is  maintained  on  the  grate,  with  speed 
of  grate  of  5  inches  per  minute.  The  grates  are  driven  by  Krehbiel 
oscillating  engines,  two  engines  being  provided  for  each  section  but 
only  one  being  in  use  at  a  time.  The  boilers  are  supported  by  ree'n- 
forced  girders  of  the  main  building  structure.  A  gallery  is  placed  in 
front  of  the  settings,  8  feet  above  the  floor,  to  facilitate  cleaning  of 
tubes.  Galleries  are  also  placed  between  the  batteries  and  on  top  of 
them.  Spaces  of  5  feet  are  provided  between  the  sides  and  rears  of  the 
batteries,  and  18  feet  8  inches  in  front.  The  furnaces  are  similar  to 
the  one  illustrated  in  Fig.  62.  The  outside  of  the  setting  is  finished 
with  red  pressed  brick.  Each  drum  is  fitted  with  a  four-inch  pop 
safety  valve.  The  blow-off  main  is  4  inches  in  diameter  and  discharges 
into  the  river.  There  are  four  blow-off  connections  to  each  boiler,  each 
being  provided  with  a  blow-off  cock  and  an  angle  valve;  three  of  the 
connections  are  fitted  to  the  mud  drum  and  the  other  to  the  super- 
heater drain. 

Chimneys.  —  One  stack  is  provided  for  each  section  of  16  boilers. 
The  shaft  is  supported  by  the  steel  work  of  the  boiler  setting,  as 
shown  in  Fig.  442,  an  arrangement  which  commends  itself  where 
space  is  limited  and  real  estate  values  are  high.  The  stacks  for  all 
units  are  259  feet  6  inches  in  height  above  the  grate  surface,  and  are 
18  feet  in  internal  diameter.  The  lining  is  of  radial  fire  brick  and 
varies  from  4  inches  to  13  inches  in  thickness.  The  steel  sections  are 
5  feet  high  and  vary  in  thickness  from  f  inch  to  J  inch.  There  are 
two  flues,  one  32  feet  long  and  the  other  63  feet,  which  enter  the  stack 
on  opposite  sides. 

Turbines.  —  The  prime  movers  are  vertical  five-stage  Curtis  turbines 
with  base  condenser  and  are  rated  at  12,000  kilowatts  each.  The  normal 
speed  is  750  r.p.m.  The  average  steam  consumption,  including  all 


'80 


STEAM  POWER  PLANT  ENGINEERING 


W 


1 


A  TYPICAL  CENTRAL  STATION  781 

auxiliaries  is  approximately  15  pounds  per  kilowatt  hour,  corresponding 
to  a  coal  consumption  of  3  pounds  per  kilowatt  hour  (Illinois  screenings, 
10,400  B.T.U.  per  pound).  Special  tests  have  shown  as  low  as  12.8 
pounds  per  kilowatt  hour,  initial  pressure  200  pounds  gauge,  150  de- 
grees superheat,  absolute  back  pressure  J  inch  of  mercury.  Each  pair 
of  units  has  a  pair  of  duplicate  pumps,  an  accumulator  and  a  storage 
tank  for  supplying  oil,  the  step-bearing  pressure  being  maintained  at 
750  pounds  per  square  inch.  When  the  accumulator  falls  below  a 
certain  point  a  motor-driven  pump  is  automatically  started. 

Generators.  —  The  generators  are  2300- volt,  25-cycle,  three-phase 
General  Electric  alternators  mounted  over  the  vertical  shaft  as  illus- 
trated in  Fig.  185.  Exciting  current  is  furnished  by 

2  50-kilowatt  motor-driven  generators. 
2  75-kilowatt  motor-driven  generators. 
2  1 50-kilowatt  motor-driven  generators. 
2  75-kilowatt  steam-driven  generators. 
2  1 50-kilowatt  steam-driven  generators. 

Part  are  held  in  reserve,  though  no  particular  units  are  maintained 
for  the  purpose.  The  high-tension  cables  lead  from  the  generator 
through  an  underground  tunnel  to  the  switch  house,  located  about 
50  feet  west  of  the  main  building.  The  oil  switches,  wattmeters  and 
other  instruments  are  located  on  the  first  floor,  while  the  bus-bars  and 
other  high-tension  connections  are  in  the  basement.  The  station  switch- 
board or  operating  gallery  in  the  main  building  is  equipped  with  only 
such  devices  as  are  necessary  for  the  control  of  the  machines,  all  other 
instruments  being  located  in  the  switch  house.  From  the  switch  house 
the  high-tension  current  is  conducted  through  oil  switches  to  the  various 
substations,  where  it  is  converted  to  direct  current  by  rotary  converters, 
or  transformed  from  25  to  60  cycles  by  motor  generator  sets. 

The  Twenty-second  Street  substation  is  located  at  the  north  end  of 
the  property  (Fig.  438).  In  this  substation  are  installed  two  motor 
generator  sets  and  one  rotary  converter,  the  latter  supplying  direct 
current  to  the  neighboring  district  and  to  the  main  station. 

Boiler  and  Turbine  Piping.  —  Immediately  below  each  boiler  section 
is  an  apartment  called  the  "  header  room,"  where  the  steam  pipes  from 
the  various  boilers  join  the  main  header,  which  increases  in  size  from 
6  inches  at  the  most  remote  boiler  to  10  inches  at  the  middle  boiler, 
and  finally  to  14  inches  where  it  leaves  the  nearest  boiler  and  passes  to 
the  turbines.  The  pipes  are  of  wrought  iron,  with  welded  flanges,  and 
are  packed  with  copper  gaskets.  The  feeder  from  each  boiler  is  6  inches 
in  diameter.  An  angle  stop  valve  and  a  check  valve  are  placed  at  the 


782  STEAM  POWER  PLANT  ENGINEERING 

boiler  nozzle  and  a  globe  valve  at  the  header.  A  motor-operated  throttle 
valve  and  strainer  are  placed  at  the  turbine,  and  a  hydraulically  oper 
ated  valve,  controlled  from  the  operating  gallery,  is  located  in  the 
header  room.  The  main  header  is  not  anchored  at  any  point,  the  entire 
weight  being  carried  by  roller  supports.  The  only  drain  in  the  header 
is  a  }-inch  bleeder  on  the  turbine  side  of  the  hydraulically  operated 
valve.  The  bleeder  is  connected  to  a  trap  which  discharges  into  either 
boiler  or  superheater  blow-off  main.  All  branches  or  feeders  are  drained 
and  discharged  into  the  superheater  blow-off. 

Condensers  and  Auxiliaries.  —  Each  unit  has  its  own  condensing 
apparatus,  feed-water  heater,  hot  well  and  feed  pumps.  The  condensers 
are  of  the  Worthington  "  base  "  type  with  25,000  sq.  ft.  of  cooling 
surface  each,  composed  of  5900-6000  1-inch  tubes  16  feet  long.  Cooling 
water  is  taken  from  the  east  slip  through  concrete  tunnels  and  is  dis- 
charged from  the  condenser  into  similar  tunnels  which  empty  into  the 
west  slip.  (See  Fig.  438.) 

The  dry  vacuum  pumps  are  of  the  rotative  type,  with  cylinders 
26  X  24,  r.p.m.  100-120,  and  are  driven  by  a  75-horse-power  Corliss 
engine. 

The  circulating  pumps  are  of  the  volute  single-stage  centrifugal  type 
and  are  mounted  on  an  extension  of  the  main  shaft  of  the  engine  driving 
the  dry  vacuum  pump.  They  are  rated  at  22,500  gallons  per  minute 
each. 

The  hot- well  pumps  are  of  the  two-stage  centrifugal  type,  driven  by 
20-horse-power  direct-current  motors. 

The  feed  pumps  are  of  the  Dean  vertical  single-cylinder  pattern  and 
are  installed  in  duplicate  for  each  unit.  The  steam  cylinders  are  24 
inches  in  diameter,  water  cylinders  14  inches  in  diameter,  and  common 
stroke  24  inches.  Feed  water  is  drawn  in  by  suction  from  the  hot  well 
at  a  temperature  of  about  100  degrees  F.  and  is  forced  through  closed 
heaters  having  3000  sq.  ft.  of  heating  surface,  and  its  temperature  is 
raised  to  180  degrees.  The  heater  receives  the  steam  exhausted  from 
the  steam-driven  auxiliaries. 

From  the  heater  this  water  is  forced  through  a  5-inch  feed  main  to  the 
different  boilers  in  the  section.  The  branches  from  main  to  boiler  drum 
are  3  inches  in  diameter  and  fitted  with  an  angle  stop  valve,  a  regrinding 
check  and  a  grate-regulating  valve.  There  is  a  5-inch  auxiliary  main 
which  supplies  cold  water  to  the  boiler  in  case  the  main  header  is  shut 
down. 


A  TYPICAL  CENTRAL  STATION 


783 


FIG.  443.     General  Plan  of  Quarry  Street  Station,  Units  1  and  2. 


784  STEAM  POWER  PLANT  ENGINEERING 

Miscellaneous.  —  The  work  is  divided  into  eight-nour  shifts.  The 
list  of  operating  men  per  unit  is  as  follows: 

In  turbine  room,  including  janitor  work 2.0 

In  oil  room 10 

Attending  water 0.5 

Fireman 1 

Fireman's  helper 1 

Conveyor  men 2 

Turbine  switchboard  gallery 0.3 

Exciter  tenders 0.2 

Switchboard  attendants   0.2 

872 

Drains  and  drips  from  the  auxiliaries  empty  into  sumps  from  which 
they  are  discharged  by  Yeoman's  bilge  pumps  into  the  discharge  tunnel. 

A  steam-driven  house  pump  is  located  in  the  basement. 

The  fire-protection  system  includes  a  220-horse-power  motor-driven 
spherical  pump  located  in  the  basement  and  a  connection  for  a  fire  tug. 

Dining  room,  reading  room,  shower  and  tub  baths  and  sleeping 
rooms  for  emergencies  are  provided  for  the  employees. 

Quarry  Street  Station.  —  Figs.  443  to  445  give  general  views  of  the 
Quarry  Street  Station,  which  is  located  directly  across  the  river  from 
the  Fisk  Street  Station.  The  two  stations  are  distinct;  a  breakdown 
in  one  would  not  affect  the  other;  nevertheless,  they  are  operated  to- 
gether. That  is,  there  is  one  chief  engineer  for  the  two,  and  the  com- 
bined station  force  of  350  men  can  be  shifted  from  one  to  the  other  as 
needed. 

The  general  layout  of  the  Quarry  Street  Station  differs  from  that  of 
the  Fisk  Street  Station  on  account  of  real-estate  limitations.  The 
boilers  are  in  two  parallel  rows  instead  of  the  equipment  for  each  unit 
extending  at  right  angles  to  the  turbine  room  as  at  Fisk. 

The  complete  station  will  contain  six  14,000-kilowatt  Curtis  turbo- 
alternators,  the  steam  for  each  unit  being  supplied  by  eight  500 
horse-power  B.  &  W.  boilers  arranged  as  shown  in  Fig.  443.  Steam  is 
generated  at  225  pounds  gauge  pressure  and  150-175  degrees  super- 
heat. The  settings  for  the  first  two  units  are  similar  to  that  illus- 
trated in  Fig.  67a,  and  the  other  similar  to  the  one  illustrated  in  Fig.  67b. 

A  novel  system  of  ventilation  enables  the  generators  to  be  operated 
continuously  at  full  load.  As  will  be  seen  from  Fig.  444  air  ducts  lead 
from  an  outside  intake  to  the  top  of  each  unit,  the  revolving  portion  of 
the  generator  being  designed  to  draw  in  a  continuous  supply  of  air  and 
discharge  it  through  openings  in  the  turbine  casing. 

Everything  is  in  duplicate,  so  that  the  chance  of  breakdown  is  remote 
There  are  two  volute  circulating  pumps,  each  driven  by  a  125  horse- 


A  TYPICAL  CENTRAL  STATION 


785 


786 


STEAM  POWER  PLANT  ENGINEERING 


power  Corliss  engine.  The  water  of  condensation  is  removed  from  the 
condensers  by  two  rotary  pumps  driven  by  Kerr  steam  turbines.  For 
each  two  units  of  turbines  and  boilers  three  horizontal  boiler  feed  pumps 
are  provided,  located  between  the  turbines  as  stated.  There  are  also 
four  step-bearing  oil  pumps,  two  oil  accumulators,  dry-air  pumps,  oil 
filters,  etc.  All  are  in  plain  view  of  the  turbine-room  operating  force. 

For  the  six  ultimate  units  five  150-kilowatt  exciters  will  be  installed, 
three  driven  by  horizontal  Curtis  steam  turbines  and  two  by  25-cycle, 
220- volt  induction  motors.  One  of  each  of  these  types  is  included  in 
the  equipment  of  the  present  plant.  In  addition  there  is  an  excitation 
storage  battery  of  70  cells  in  the  basement.  Furthermore,  in  an  emer- 
gency the  split-pole  rotary  converter  of  the  substation  could  be  used 
for  excitation. 

Northwest  Stations.  —  Two  new  stations  of  120,000  kilowatt  rated 
capacity  each  are  to  be  installed  on  the  north  branch  of  the  Chicago 
River  near  Roscoe  Street  and  California  Avenue.  Each  station  is  to  be 
equipped  with  six  20,000-kilowatt  Curtis  turbo-generators,  2300  volt, 
25  cycle,  three  phase,  750  r.p.m.,  similar  to  those  installed  at  Quarry 
Street.  The  first  two  units  for  one  station  are  now  in  course  of  erection. 
Each  unit  is  to  be  supplied  with  steam  at  250  pounds  gauge  pressure  and 
150  degrees  superheat  from  ten  500-horse-power  B.  &  W.  boilers. 
The  boiler  settings  are  to  be  similar  to  the  one  illustrated  in  Fig.  67b. 
The  general  layout  will  be  the  same  as  at  Fisk,  the  boiler  lanes  extend- 
ing at  right  angles  to  the  turbine  room.  There  will  be  one  chimney 
17  feet  inside  diameter  and  250  feet  in  height  for  every  ten  boilers. 

The  present  capacity  (July,  1910)  of  the  Commonwealth  Edison 
Company  is  about  240,000  kilowatts,  divided  as  follows: 


Units. 

Present 
Capacity. 

Ultimate 
Capacity. 

Fisk  

10-12  000 

120  000 

120  000* 

Quarry 

6-14  000 

84  000 

84  000 

Northwest  No.  1  

6-20  000 

120  000 

Northwest  No.  2  

6-20  000 

120  000 

Miscellaneous  Plants  

36,000 

36,000 

240,000 

480,000 

*  Space  is  available  for  four  additional  units,  but  no  increase  is  contemplated  at  present. 
COMPARATIVE    BOILER    ROOM    AND    ENGINE    ROOM    AREAS. 


Fisk. 

Quarry. 

Northwest. 

Boiler  room   sq.  ft.  per  kw     

0  51 

0  44 

Engine  room,  sq.  ft.  per  kw  

0  24 

0.18 

0  15 

Total  area  sq    it   per  kw 

0  75 

0  59 

A  TYPICAL  CENTRAL  STATION 


787 


FIG.  445.     Sectional  Elevation,  Quarry  Street  Station. 


CHAPTER  XXI. 

A  TYPICAL  ISOLATED  STATION.  — MANUFACTURING  PLANT. 

The  West  Albany  Power  Station.* 

THIS  modern  and  well-designed  station  was  erected  (1904)  in  place 
of  an  old  and  practically  worn-out  plant,  and  is  an  example  of  what 
may  be  accomplished  with  a  limited  appropriation. 

The  power  generated  is  used  about  the  repair  and  construction  shops 
of  the  New  York  Central  Railroad  Company  at  West  Albany,  N.Y., 
for  lighting  and  general  power  purposes,  and  steam  is  used  for  heating. 
The  equipment  includes  two  direct-connected  600-kilowatt  generators 
and  two  direct-acting  air  compressors. 

General. — The  plant  (Fig.  446)  is  housed  in  a  brick  building  92 
feet  8  inches  wide  by  113  feet  4  inches  long,  divided  by  a  fire  wall  into 
an  engine  and  a  boiler  room,  the  former  being  46  feet  wide. 

The  main  units  consist  of  two  cross-compound  non-condensing 
Ball  &  Wood  engines  and  two  Chicago  Pneumatic  Tool  Company's  air 
compressors.  Steam  is  supplied  by  two  batteries  of  horizontal  water- 
tube  boilers,  comprising  four  boilers  of  500  horse  power  each,  made  by 
the  Franklin  Boiler  Works  Company.  These  boilers  have  a  guar- 
anteed normal  rating  of  528  horse  power  under  a  natural  draft  not 
exceeding  T7^  inch  of  water  measured  in  the  flue  adjacent  to  the  boiler, 
and  an  evaporation  of  not  less  than  9  pounds  of  water  per  pound  of 
coal  from  and  at  212  degrees  F.  They  contain  5280  square  feet  of 
heating  surface  and  84  square  feet  of  grate  surface,  a  ratio  of  63  to  1. 
There  are  300  3^-inch  tubes  13  feet  long,  surmounted  by  two  48-inch 
steam  drums.  The  steam  drums  are  connected  by  a  5-inch  balance 
pipe  on  which  two  safety  valves  are  mounted. 

The  boilers  are  suspended  from  wrought-iron  and  steel  frames 
entirely  independent  of  the  brickwork,  eliminating  strains  due  to 
expansion  and  contraction. 

The  smoke  flues  run  over  the  boilers  to  the  stack  placed  between  the 
two  batteries,  as  shown  in  Figs.  447  and  448.  Dampers  are  placed  in 
both  flues  and  hoods,  the  former  operated  by  automatic  regulators  and 
the  latter  by  hand,  although  the  flue  dampers  may  be  readily  discon- 
nected and  also  operated  by  hand. 

*  See  Power,  November,  1904. 

788 


A  TYPICAL  ISOLATED  STATION 


789 


The  steam  nozzles  are  connected  to  dry  pipes  in  each  drum  and  to 
elbows  outside,  between  which  is  a  cross-over  pipe  carrying  the  main 
10-inch  valve,  which  is  of  the  automatic,  non-return,  hand  and 
emergency  angle  type.  Between  the  angle  stop  valve  and  the  main 
near  the  elbow  is  a  gate  valve.  Long  bends  in  the  10-inch  pipe 
between  stop  valve  and  main  are  relied  on  for  the  necessary  flexibility, 
and  the  main  header  is  anchored  to  the  stack  near  its  middle  point,  as 
seen  in  Fig.  450,  and  its  ends  allowed  free  movement. 


A  specially  designed  roller  suspension  is  used  on  the  larger  pipes,  and 
where  pipes  are  swung  from  the  floor  beams  a  turn-buckle  is  provided 
in  each  suspension  rod  for  adjustment. 


790 


STEAM  POWER   PLANT  ENGINEERING 


A  TYPICAL  ISOLATED  STATION 


791' 


792  STEAM  POWER  PLANT  ENGINEERING 

The  steam  piping  slopes  so  as  to  drain  to  the  drop  legs,  the  drips 
being  returned  to  the  boiler  by  the  Holly  return  system.  Steam  is 
led  to  the  two  large  engines  through  9-inch  pipes,  as  shown  in 
Figs.  449  and  450,  a  separator  being  placed  just  above  each  engine 
throttle. 

Engines.  —  The  engines  have  cylinders  21  and  41  inches  in  diameter 
and  a  stroke  of  30  inches.  They  develop  the  normal  load  of  900 
horse  power,  cutting  off  at  37  per  cent  of  the  stroke  in  each  cylinder 
with  steam  at  175  pounds  pressure  and  running  at  120  revolutions  per 
minute.  The  guaranteed  steam  consumption  at  this  load  is  19  pounds 
per  indicated  horse  power.  The  generators,  made  by  the  General 
Electric  Company,  are  mounted  between  the  cylinders  and  supply 
three-phase  alternating  current,  60  cycles  per  second  at  480  volts. 

Speed  regulation  is  sufficiently  close  to  allow  parallel  operation  of 
the  generators,  so  that  current  is  delivered  to  one  bus-bar  on  the  main 
switchboard  in  the  engine  room.  Exhaust  steam  from  the  high-pres- 
sure cylinder  is  led  to  a  vertical  receiver  below  the  engine-room  floor, 
as  shown  in  Figs.  449  and  450,  and  from  there  to  the  low-pressure 
cylinder.  The  low-pressure  cylinders  exhaust  through  16-inch  pipes 
into  a  24-inch  main  extending  nearly  the  whole  length  of  the  power 
house.  This  main  also  receives  the  exhausts  from  the  exciter  units, 
air  compressors,  and  auxiliaries. 

Exciters.  —  The  two  steam-driven  exciter  units  consist  of  35-kilo- 
watt  dynamos  direct  connected  to  60-horse-power  Woodbury  engines 
made  by  the  A.  D.  Granger  Company.  The  engines  are  7x12  single- 
cylinder  automatic,  and  at  300  revolutions  per  minute  and  a  normal 
cut-off  of  20  per  cent  develop  58  horse  power. 

Besides  these  steam-driven  units  is  a  100-kilowatt  General  Electric 
motor-generator.  Fig.  451  gives  a  diagrammatic  outline  of  the  switch- 
board connections. 

The  two  air  compressors  were  made  by  the  Chicago  Pneumatic 
Tool  Company  and  are  of  the  cross-compound,  two-stage  type.  The 
steam  cylinders  are  16  and  27  inches  in  diameter,  the  air  cylinders  are 
14  and  24  inches  in  diameter,  the  common  stroke  18  inches,  and  the 
capacity  1225  cubic  feet  of  free  air  per  minute. 

Steam  Piping.  —  Steam  is  supplied  to  the  exciter  engines  and  air 
compressors  through  an  auxiliary  steam  line  which  runs  nearly  the 
length  of  the  power  house  in  the  basement  beneath  the  engine  room 
as  shown  in  Figs.  449  and  450.  It  is  anchored  near  each  end,  and  the 
long  radius  vertical  bend  near  its  middle  point  allows  for  expansion. 
It  draws  steam  from  the  main  steam  pipe  in  the  boiler  room  through 
the  vertical  pipe  sho'wn  in  Figs.  446  and  447.  This  pipe  is  taken  out  of 


794 


STEAM  POWER  PLANT  ENGINEERING 


the  main  at  the  top,  makes  a  semicircular  bend  and  passes  through 
the  wall  between  the  engine  and  boiler  rooms  and  enters  the  auxiliary 
main  at  the  top. 

Connections  are  made  from  the  auxiliary  main  to  the  low-pressure 
cylinders  of  the  900-horse-power  compound  engines  so  that  they  may 


be  supplied  with  high-pressure  steam  should  an  occasion  arise  which 
would  demand  it. 

The  use  of  an  auxiliary  main  may  at  first  seem  extravagant,  but  a 
careful  study  of  the  piping  arrangement  will  show  that  it  does  not 
entail  the  use  of  more  pipe  or  fittings  and  is  a  very  desirable  arrange- 


A  TYPICAL  ISOLATED  STATION 


795 


& 

H 

i 

s  ^n 

3 

!! 

I 

0 

5 


796  STEAM   POWER   PLANT  ENGINEERING 

ment,  particularly  since  it  leaves  the  lines  to  the  main  engines  intact 
and  direct. 

Connections  to  the  other  engines  are  direct,  practically  straight,  and 
simple,  and  it  is  easy  to  make  connections  in  both  the  steam  and 
exhaust  lines  with  any  other  apparatus  which  may  be  added  in  the 
future.  In  fact,  the  simplicity  and  yet  completeness  of  the  piping 
scheme  is  perhaps  one  feature  of  the  plant. 

Connection  is  made  at  the  left  of  the  auxiliary  steam  line,  as  shown 
in  Figs.  449  and  450,  to  a  steam  supply  for  the  shops  at  high  pressure, 
or  through  a  reducing  valve  at  low  pressure  for  heating. 

The  exhaust  line,  which  runs  nearly  the  length  of  the  building  in  the 
basement  and  receives  exhaust  steam  from  all  sources,  shunts  part  of 
it  through  an  exhaust  muffler  and  oil  separator  to  the  shops  for 
heating  and  the  rest  of  it  through  an  open  heater  and  out  of  the 
exhaust  pipe  through  the  roof  (Figs.  448  to  450).  Either  the  muffler 
or  the  heater  may  be  by-passed  and  the  exhaust  caused  to  flow 
directly  to  the  heating  system  or  the  atmosphere. 

Expansion  in  the  exhaust  line  is  provided  for  by  two  copper  expansion 
joints,  and  an  approved  exhaust  head  is  placed  on  the  line  extending 
above  the  roof. 

All  high-pressure  steam  and  water  piping  is  of  "  special  full  weight  " 
lap- welded  pipe  and  all  low-pressure  pipe  of  "  standard  "  weight.  Steam 
and  hot-water  pipes  are  covered  with  the  best  non-conducting  cover- 
ing. The  high-pressure  joints  are  made  up  with  Merworth  copper 
gaskets. 

Two  duplex  Worthington  feed  pumps,  made  by  the  International 
Steam  Pump  Company,  supply  the  boiler  feed  through  the  lines  shown 
in  Figs.  446,  447,  and  450.  The  source  of  water  supply  is  the  city 
mains;  hence  the  use  of  the  meters. 

Figs.  447  and  450  show  the  arrangement  of  the  blow-off  piping. 
There  is  a  system  of  piping  for  fire  protection  which  is  connected  to  the 
city  supply  and  runs  in  gradually  reduced  sizes  as  shown  in  Fig.  450. 
This  system  is  laid  out  so  as  to  drain  itself  through  an  arrangement  of 
pipes  discharging  into  the  sewer. 

Chimney.  —  A  sectional  view  of  the  stack  which  furnishes  draft  for 
all  of  the  boilers  is  shown  in  Fig.  446.  It  is  10  feet  in  diameter  on  the 
inside  and  165  feet  high,  and  designed  to  provide  a  draft  of  50  per  cent 
above  the  normal  rating  of  the  boilers,  equivalent  to  1.1  inches  of 
water  with  the  temperature  of  the  heated  gases  not  over  500  degrees  F. 
It  is  made  of  radial  bricks  and  has  a  baffle  wall  through  the  center 
for  a  short  distance  from  the  base  to  prevent  the  gases  from  the  flues 
on  each  side  from  impinging  and  causing  eddies.  The  wall  is  20  J-  inches 


A  TYPICAL  ISOLATED  STATION  797 

thick  at  the  base  and  7J  at  the  top.  The  stack  was  built  by  M.  W. 
Kellogg  &  Co. 

The  appropriation  for  the  power  plant  did  not  allow  the  installation 
of  a  complete  coal  and  ash-handling  apparatus,  but  provision  was 
made  for  such  an  installation  in  the  future,  and  for  the  present  the 
apparatus  shown  in  Fig.  448  is  used.  The  ashes  fall  from  the  grate 
to  the  ashpit  (Fig.  447),  and  are  raked  out  through  the  door  into  a 
barrow  in  the  adjacent  passageway,  from  which  they  are  discharged 
into  the  vertical  conveyor  (Fig.  448)  and  transferred  to  a  hopper  and 
chute  suspended  from  the  roof  beams  and  from  there  through  a  spout 
to  cars  outside.  The  coal  is  delivered  over  this  track  into  the  coal  bin 
and  carried  to  the  boiler  fronts  in  barrows.  It  is  seen  that  the 
arrangement  allows  for  the  installation  of  a  complete  coal  and  ash- 
handling  apparatus  of  the  continuous  type,  and  indeed  one  of  the 
features  of  design  is  the  provision  for  the  installation  of  modern  appli- 
ances in  the  future  which  are  not  possible  at  present. 

The  general  arrangement  of  the  boilers,  engines,  and  piping  is  such 
that  additional  units  of  each  may  be  installed  by  enlarging  the  building 
at  one  end  without  disarrangement  of  the  apparatus  already  in  place. 


APPENDIX   A. 

GENERAL  BIBLIOGRAPHY  —  POWER  PLANT  ENGINEERING. 
DESCRIPTION  OF  POWER  PLANTS. 

GAS  DRIVEN.  Central  Stations. 

A  Power  Plant  with  500-Horse-Power  Gas  Engines, 

Engineering  Record 51:178  Feb.  18,  1905 

Gas   and   Electric   Power   Plant   for   a   Railroad 

Terminal,    Engineer   (United   States) 42: 233  April  1,  1905 

Gas-Driven  Electric  Power  Station  at  Bridgewater, 

England,  Power 25:  273  May,  1905 

Power  Installation  at  Isle  of  Elba  for  Blast  Furnace 

Gas,  Power 24:  460,  549 

Gas  Power  Plant  at  Haysam  Harbor,  Power 25:  212  April,  1905 

Lacka wanna  Steel  Company's  Power  Plant,  Power     23:  663  Dec.,  1903 

Gas  Power  Plant  for  the  Morecambe  Bay  Harbor 

Works,  Engineering  Record 50:  467  Oct.  15,  1904 

Power  and  Mining  Machinery  Company's  Gas  and 

Electric  Plant,  Electrical  World 44:  442  Sept.  10,  1904 

1020  Dec.  10,  1904 

Engineering  Record 50:  652  Dec.  3,  1904 

Engineer  (United  States) 41 :  829  Dec.  15,  1904 

Some  Features  of  the  Warren  Gas  Power  Plant, 

Electrical  Journal 3:  205  April,  1906 

Proceedings  of  Engineers'  Society  of  Western 

Pennsylvania 22:  290  July,  1906 

Trials  of  Suction  Gas  Producer  Plants,  Mechanical 

Engineering 16:  707  Nov.  11,  1905 

Power 26:287  May,  1906 

Engineering  News 55:  538  May  17,  1906 

Engineering 82:  205  Aug.  10,  1906 

A  Year's  Experience  with  Gas  Engines,  Power 27:  831  Dec.,  1907 

The  Producer  Gas  Plant,  Railway  and  Engineering 

Review 47:8  Jan.  5,  1907 

Engineer  (United  States) 44  Aug.  15,  1907 

Electrician  (London) 58:  642  Feb.  8,  1907 

Electrical  Review  (London) 60:201  Feb.  1,  1907 

Journal  Association  of  Engineering  Societies. .  .     38: 14  Jan.,  1907 

Central  Station June,  1907 

HYDRAULIC. 

Animas  Power  Development  in  Colorado,  Engineer- 
ing Record 50:519  Oct.  29,  1904 

Engineering  News 50:616  June  3,  1905 

55:  1  Jan.  4,  1906 

798 


APPENDIX  A 


799 


HYDRAULIC  — Continued. 
American     River    Electric     Company's     System, 

American  Electrician 17:1  Jan.,  1905 

Water  Power  on  the  Apple  River,  Engineering  News  54:  374  Oct.  12,  1905 

Engineering  Record 52:  431  Oct.  14,  1905 

Atlanta,  Georgia,  Engineering  Record 49:  504  April  23,  1904 

Berkshire  Hydro-Electric  Plant  (H.  S.  Knowlton), 

Electrical  Age 36:  107  Feb.,  1906 

Bay  Counties  Power  Company,  California  (I.  D. 

Galloway),  Engineering  News. 46:  230  Oct.  3,  1901 

British  Columbia  Mines  Power  and  Light,  Western 

Electrician 32:  321  April  25,  1903 

Catawba   River   Power  Development   near  Rock 

Hill,  South  Carolina,  Engineering  Record 50:  114  July  23,  1904 

129  July  30,  1904 

Chelton  Hills,  Pennsylvania,  Electrical  World 35:  487  March  31,  1900 

Champ  Generating  Station,  American  Electrician ...  17:  65  Feb.,  1905 

Chicago  Drainage  Canal,  Engineering  News 55:  52  Jan.  18,  1906 

Columbus  Power  Company,  Georgia,  Engineering 

Record 49:  64  Jan.  16,  1904 

Chattanooga    and    Tennessee     Power    Company, 

Engineering  Record 52:  576  Nov.  4,  1905 

Chittenden  Power  Company,   Rutland,   Vermont, 

Engineering  Record 52:  653  Dec.  9,  1905 

Cornell  University  Power  Plant,  Engineering  Record  51 :  562  May  20,  1905 
Central    California    Electric    Company,    Engineer 

(United  States)  40: 160  Feb.  16,  1903 

Cudahy  Packing  Company,  Engineer  (United  States)  38:  82  March  1,  1901 

Western  Electrician 28:  297  May  4,  1901 

Central  Lard  Company,  Engineering  Record 46:  122  Aug.  9,  1902 

Carpenter  Steel  Works,  Engineering  Record 46:  220  Sept.  6,  1902 

Deere   and   Company,    Moline,    Illinois,   Engineer 

(United  States)  40:  617  Aug.  15,  1903 

Deering  Harvester  Company,  American  Electrician  12:  381  Aug.,  1900 

Dixon  Crucible  Company,  American  Electrician.  ...  13:  517  Nov.,  1901 
Edison    Portland    Cement     Company,     Electrical 

World 42:  1051  Dec.  26,  1903 

Ford    Plate    Glass    Company,     Engineer    (United 

States) 38:  107  April  1,  1901 

Henry  Heide  Candy  Company,  Engineering  Record  51 :  269  May  20,  1905 
Ingersoll-Sargent      Drill      Company,       American 

Electrician 17:  237  May,  1905 

Power 25:  323  June,  1905 

Littleton  Creamery,  Denver,  Engineering  Record .  .  .  50:  159  Aug.  6,  1904 
Laray  Mills,   Gastonia,  North  Carolina,   Engineer 

(United  States) 40:  222  March  16,  1903 

Ladew  Factory,  Power 25:  468  Aug.,  1905 

Patton    Paint    Company,    Newark,    New   Jersey, 

Electrical  World. 42:  454  Sept.  12,  1903 

Engineer  (United  States) 40:  798  Oct.  15,  1903 


800 


STEAM  POWER  PLANT  ENGINEERING 


HYDRAULIC  — Continued. 

Modern  Malt  House,  Western  Electrician 37:  101  Aug.  5,  1905 

John  Mehl  Factory,  Jersey  City,  Engineer  (United 

States) 38:  245  July  1,  1901 

Manchester,     New     Hampshire,    Mills,     Electrical 

World 41:  269  Feb.  14,  1903 

Monarch  Mills,  Union,  South  Carolina,  Engineering 

Record 50:  380  Sept.  24,  1904 

Lake  Superior  Company,  St.  Mary's  Falls,  Michigan, 

Engineering  News 40:  68  Aug.,  1898 

Montmorency     Falls,     Canada,     American     Elec- 
trician   12: 553  Dec.,  1900 

Missouri  River  Power  Company,  American  Elec- 
trician   14:323  July,  1902 

Montreal  Power  Plant,  Engineer  (London) 101 :  130  Feb.  9,  1906 

Monterey   Gas   and   Electric   Company,    Engineer 

(United  States) 42:  87  Jan.  16,  1905 

Montgomery,  Alabama,  Engineering  News 46:  418  Dec.  5,  1901 

Massena,   New   York,   Power  Plant,    Engineering 

News 45:130  Feb.  21,  1901 

Engineering  Record 41 :  2  Jan.  6,  1900 

122  Feb.  10,  1900 

Western  Electrician 22:  22  Jan.  8,  1898 

Mexican  Light  and  Power  Company,  Engineering 

Record 51:  575  May  20,  1905 

Power  Plant  of  the  Milano-Gallarate-Porto  Ceresio 

Railway,  Engineer  (United  States) 40:  25  Jan.  1,  1903 

American  Electrician 14:  556  Dec.,  1902 

New   Water  Power  Transmission   Plants    (C.   L. 

Fitch),  Cassier's 19:  243  Feb.,  1901 

North     Mountain      Power     Company,     Engineer 

(United  States) 43:  123  Feb.  1,  1906 

Engineering  Record 53:  27  Jan.  6,  1906 

NewMilford,  Connecticut,  Engineering  Record 49:  187  Feb.  13,  1904 

230  Feb.  20,  1904 
Northern  California  Power  Company,  Engineering 

Record 50:  506  Oct.  29,  1904 

Niagara  (J.  E.  Woodbridge),  American  Electrician.  12:  1  Jan.,  1900 
Transactions  American  Society  of   Mechanical 

Engineers 19:  839           June,  1898 

Extension  of  Niagara  Falls  Power  Plant,  American 

Machinist 21 :  463  June  23,  1898 

Western  Electrician 22:137  March  5,  1898 

Niagara  Falls  Power  (H.W.  Buck),  Cassier's 20:  1  May,  1901 

179  Jan.,  1902 

25:  103  Dec.,  1903 

Engineering 74:  637  Nov.  12,  1902 

Engineering  News 48:9                July  3,  1902 

250  Oct.  2,  1902 
490  Dec.  11,  1902 


APPENDIX  A 

HYDRAULIC  —  Continued. 

Power  Plant  of  the  Niagara  Falls  (Coleman  Sellers), 


801 


Engineering  

67:91 

Jan.  20,  1899 

128 

Jan.  27,  1899 

160 

Feb.  3,  1899 

Canadian  District  Plant  (C.  B.  Smith),  Engineering 

Magazine  

28:  727 

Feb.,  1905 

New  Wheel   Pit   Niagara  Falls   Power  Company 

(O.  E.  Dunlap),  Engineering  News  

43:229 

April  5,  1900 

Engineering  Record  

43:  150 

Feb.  16,  1901 

Power  of  Lower  Niagara  (O.  E.  Dunlap),  Western 

Electrician  

32:360 

June  18,  1898 

Niagara    Falls    (L.    B.    Stillwell),    Western   Elec- 

trician   

29:  191 

Sept.  21,  1900 

Progress    on    Power   House    at    Niagara    (O.    E. 

Dunlap),  Western  Electrician  

29:234 

Oct.  12,  1901 

Niagara  Power  Plant  of  Ontario,  Engineering  News 

54:475 

Nov.  9,  1905 

561 

Nov.  30,  1905 

Power    Station    at    Neuchatel,    American    Elec- 

trician   

16:493 

Oct.,  1904 

Norwegian  Hydro-Electric  Plant,  Electrical  World.. 

46:  135 

July  22,  1905 

Western  Electrician  

37:63 

July  22,  1905 

Hydro-Electric  Power  Plant  of  Christiania,  Norway, 

Engineer  (United  States)  

40:227 

March  16,  1903 

Works  of  the  Ontario  Power  Company,  Engineering 

Record  

50:420 

Oct.  8,  1904 

460 

Oct.  15,  1904 

480 

Oct.  22,  1904 

29 

Oct.  29,  1904 

Oliver  Plow  Company,  Engineer  (United  States)  .  .  . 

42:363 

June  1,  1905 

50:420 

Oct.  8,  1904 

460 

Oct.  15,  1904 

480 

Oct.  22,  1904 

504 

Oct.  29,  1904 

Ottawa-Ontario   Power  Company,    Western  Elec- 

trician   

27:85 

Aug.  11,  1900 

Port  Huron  Light  and  Power  Company,  Engineering 

Record  

49:458 

April  9,  1904 

The  Puyallup  River  Water  Power  Development, 

Engineering  News  

52:273 

Sept.  29,  1904 

Engineering  Record  

50:399 

Oct.  1,  1904 

Saut  Mortier  Transmission  Plant,  American  Elec- 

trician   

17:  457 

Sept.,  1905 

The  Sill  Hydraulic  Power  Plant  near  Innsbruck, 

Engineering  Record  

52:13 

July  1,  1905 

Standard    and    Bay    Company    Plant,    Engineer 

(United  States)  

40:46 

Jan.  1,  1903 

Snoqualmie  Falls,  American  Electrician  

13:497 

Oct.,  1901 

Power  

19:1 

Oct.,  1899 

44:  398 

Dec.  13,  1900 

802 


STEAM  POWER  PLANT  ENGINEERING 


HYDRAULIC —  Continued. 

Swiss  Combined  Water  Power  and  Gas  Engine 

Plant,  American  Electrician 15:  113  March,  1903 

St.  Croix  Power  Plant,  Engineering  Magazine 20:  954  Feb.,  1901 

Sault  Sainte  Marie  Power  Plant  (H.  Von  Schon), 

Engineering  Magazine 24:  273  Nov.,  1902 

Combined  Hydraulic  and  Steam  Plant,  Stuyvesant 

Falls,  New  York,  Engineering  Record 43:  3  Jan.  5,  1901 

Sterling  Hydraulic  Company,  Engineering  Record..  52:  688  Dec.  16,  1905 

St.   Maurice-Lausanne,  Switzerland,  Western  Elec- 
trician   33:477  Dec.  20,  1903 

Sewanee  Falls  Power  Plant,  Engineering  Record. . .  53:  44  Jan.  13,  1906 

Shawinigan  Falls,  Canada,  Cassier's 26:  202  July,  1904 

Engineering  Record 41:391  April  29,  1900 

Hydro-Electric    Development    at    Turner's    Falls, 

Electrical  World 46:  263  Aug.  12,  1905 

STEAM  ENGINE  PLANTS. 

Aurora,  Elgin  and  Chicago  Railway,  Street  Railway 


Journal  

19:  143 

Feb.,  1902 

20:574 

Oct.  4,  1902 

Augusta  Railway  Power  Station,   Street  Railway 

Journal  

21:24 

Jan.  3,  1903 

Brooklyn     Rapid     Transit     Company,     Electrical 

World  

46:519 

Sept.  23,  1905 

Engineering  Record  

52:254 

Sept.  23,  1905 

Street  Railway  Journal  

21:256 

Feb.  14,  1903 

26:432 

Sept.  23,  1905 

Street  Railway  Review  

14:332 

May  20,  1904 

Berlin     Elevated     and     Underground     Railway, 

Engineering  Record  

45:389 

April  26,  1902 

Boston   Elevated   Power   Station,   Street  Railway 

Journal  

17:  253 

March,  1901 

20:  118 

July  26,  1902 

Boston  and  Manhattan  Railway,  Street  Railway 

Journal  

20:  554 

Oct.  4,  1902 

Some    British    Central    Electric    Power    Stations 

(H.  F.  Parshall),  Cassier's  

26:233 

June,  1904 

Breslau,  Germany,  Power  

24:  742 

Dec.,  1904 

Brigham    City,    Utah,    Light    Plant,    Engineering 

News  

44:235 

Sept.  7,  1905 

Boston  Elevated,  Lincoln  Wharf  Station,  Engineer 

(United  States)  

40:4 

Jan.  1,  1903 

Berlin  Electric  Railway,  Engineer  (United  States)  .  . 

40:16 

Jan.  1,  1903 

Brussels    Electric    Power   Station   at   Anderlecht, 

Engineer  (United  States)  

41:555 

Aug.  15,  1904 

Oct.  14,  1905 

Boston  and  Worcester  Street  Railway,  Engineer 

(United  States) 40:  507 

Engineering  Record 52 :  438 


July  1,  1903 


APPENDIX  A 


803 


STEAM  ENGINE  PLANTS — Continued. 

Brooklyn  Navy  Yard,  Power 20:  21  Dec.,  1900 

Berkshire  Street  Railway  Company,  Street  Railway 

Review 12 :  813  Nov.  20,  1902 

American  Electrician 15:  387  Aug.,  1903 

Engineering  Record 46:  74  July  26,  1902 

Brussels  Electric  Tramways,  Power 25:  1  Jan.,  1905 

Bloomington  and  Normal  Railway,  Street  Railway 

Journal 25:  934  May  27,  1905 

Cranford,  New  Jersey,  Engineering  Record 49:  355  March  19,  1904 

Consolidated      Traction      Company,       Pittsburg, 

Street    Railway    Journal 15:127  March,  1899 

Street  Railway  Review 9:  135  Feb.  15,  1899 

Cleveland     Electric     Railway     Company,     Street 

Railway  Journal 15:199  April,  1899 

267  May,  1899 

17:  655  June,  1901 

19:500  April,  1902 

23:  162  Jan.  30,  1904 

Conneaut  and  Erie,  Street  Railway  Journal 23:  195  Feb.  6,  1904 

Canton  and  Akron,  Street  Railway  Journal 23 :  805  May  28,  1904 

Cleveland,  Painesville,  and  Ashtabula,  Street  Rail- 
way Journal 24:  93  July  16,  1904 

Chicago     and    Western    Indiana,     Western    Elec- 
trician    37:  425  Dec.  2,  1905 

Cincinnati,  Georgetown  and  Portsmouth  Railway, 

Engineer  (United  States) 40:  353  May  15,  1903 

Clyde  Valley  Electric  Power  Company,  Engineer 

(United  States) 42:  524  Aug.  1,  1905 

Capitol  Traction  Company,  Washington,  District 

of  Columbia,  Power 19:  1  Feb.,  1899 

Engineering  Record 39:  99  Dec.  31,  1898 

Street  Railway  Journal 15:  9  Jan.,  1899 

Citizens'  Light  and  Power  Company,  Power 22:  1  Feb.,  1902 

Chicago  Electric  Traction  Company,  Street  Rail- 
way Review 8:  184  March  15,  1898 

Engineering  News 39:  2  Jan.  6,  1898 

Chattanooga    Electric    Company,    Street    Railway 

Review 15:  136  March  15,  1905 

Carville    (England)   Power  Plant,   Street  Railway 

Journal 21 :  902  June  20,  1903 

Dublin  United  Tramways,  Power 19:  1  Jan.,  1899 

Detroit,  Ypsilanti   and  Ann  Arbor  Railway,  Street 

Railway  Review 10:  5  Jan.  15,  1900 

Des  Moines,  Iowa,  Engineering  Record 46:  25  July  12,  1902 

Street  Railway  Journal 22:  54  July  11,  1903 

Denver  Tramway  System,  Street  Railway  Journal .  .  22:  683  Oct.  10,  1903 
Dayton     and     Muncie    Electric    Railway,     Street 

Railway  Journal 26:  979  Dec.  2,  1905 

Detroit  United  Railway,  Street  Railway  Journal...  20:  445  Oct.  4,  1902 


804 


STEAM  POWER  PLANT  ENGINEERING 


STEAM  ENGINE  PLANTS  —  Continued. 

Everett  Railway  and  Electric  Company,  Engineer 

(United  States) 40:  839  Nov.  16,  1903 

Exeter,  Hampton  and  Amesbury  Street  Railway, 

Street  Railway  Review 10:  630  Nov.  15,  1900 

Edison     Electric     Company,     Atlantic     Avenue 
Station,  Boston,  Transactions  American  Society  of 

Mechanical  Engineers 23:  569  May,  1902 

Glasgow  Municipal  Tramways,  Power 22:  1  Aug.,  1902 

Street  Railway  Journal 15:  247  April,  1899 

17:  625  June,  1901 

Georgetown,  Colorado,  Western  Electrician 27:  282  Nov.  3,  1900 

Grand  Rapids,  Holland  and  Lake  Michigan  Rail- 
way, Engineer  (United  States) 39:  179  March  15,  1902 

Hanover  (Germany)  Power  Station,  Power 25:  583  Oct.,  1905 

Western  Electrician 38:  55  Jan.  20,  1906 

Hartford    and    Springfield    Railway,    Engineering 

Record 45:  485  May  24,  1902 

Halifax,     England,     Tramway,     Street     Railway 

Journal 20:  293  Sept.  6,  1902 

Harrisburg,  Pennsylvania,  Street  Railway  Journal.  27:  61  Jan.  6,  1906 

Hazleton,  Pennsylvania,  Street  Railway  Journal ...  21 :  350  March  7,  1903 

Hebron,  Ohio,  Street  Railway  Journal 22:  147  Aug.  1,  1903 

Interbo rough  Rapid  Transit  Company,  American 

Electrician 16:501  Oct.,  1904 

Engineering  Record 51 :  345  March  25,  1905 

50:  384  Oct.  1,  1904 

424  Oct.  8,  1904 

456  Oct.  15,  1904 

490  Oct.  22,  1904 

510  Oct.  29,  1904 

541  Nov.  5,  1904 

Western  Electrician 35:211  Sept.  17,  1904 

Power 24:511  Sept.,  1904 

Indianapolis   and  Cincinnati   Traction  Company, 

Engineering  Record 51 :  329  March  18,  1905 

Street  Railway  Journal 25:  502  March  18,  1905 

Interurban     Railway    and    Terminal    Company, 

Cincinnati,   Engineer  (United  States) 41:  251  April  1,  1904 

Independent  Electric  Light  and  Power  Company, 

San  Francisco,  Power 22 :  1  March,  1902 

Johnstown,    Pennsylvania,    Power   Station,    Street 

Railway  Review 11 :  426  July  15,  1901 

Johnston  and  Gloversville  Railway,  Street  Railway 

Review 22:  308  Aug.  22,  1903 

Kankakee  Power  and  Light  Plant,  Western  Elec- 
trician   26: 127  March  3,  1900 

Kingsbridge  Power  Station,  Engineering  Record.  .  .  50:  10  July  2,  1904 

Engineering  News 43: 189  March  22,  1900 

Central  Station  at  Kien,  Engineer  (United  States) .  .  42 :  337  May  15,  1905 


APPENDIX  A 


805 


STEAM  ENGINE  PLANTS — Continued. 

Kaw   River  Power  Station,    Kansas   City,   Street 

Railway  Review 10:  560  Oct.  15,  1900 

Louisville  and  Nashville  Railway  Company,  South 

Louisville,  Kentucky,  Engineer  (United  States) . .     42:  499  Aug.  1,  1903 

Louisville    Railway   Company,   Engineer    (United 

States) 42:  511  Aug.  1,  1905 

Street  Railway  Review 9:439  July  15,  1899 

LaBella  Power  Plant,  Goldfield,  Colorado,  Western 

Electrician 25:  267  Nov.  4,  1899 

Lisbon    Railway    Power    Station,    Street   Railway 

Journal 17:  293  March,  1901 

Liverpool      Electric      Railway,      Street     Railway 

Journal 17:  390  April,  1901 

London  Power  Stations,  Street  Railway  Journal 18:  287  Sept.,  1901 

Lackawanna     and     Wyoming     Valley     Railroad, 

Engineering  Record 47:624  June  13,  1903 

Street  Railway  Journal 21:868  June  13,  1903 

London  County  Council  Tramway,  Street  Railway 

Journal 21:  821  June  6,  1903 

Leicester  Power  Station,  Street  Railway  Journal 23 :  832  June  4,  1904 

Manchester  Traction  Company   (W.   V.  Batson), 

Engineering  Record 47:  107  Jan.  24,  1903 

Street  Railway  Journal 20:300  Sept.  6,  1902 

25:817  May  6,  1905 

Muskegon,  Michigan,  Light  and  Traction  Company, 

Engineering  Record 48:  452  Oct.  17,  1903 

Metropolitan  West  Side  Elevated,  Chicago,  Street 

Railway  Journal 15:581  Sept.,  1899 

709 
MeKeesport  and  Connellsville  Railway,  Engineering 

Record 48:  264  Sept.  5,  1903 

Manx    Railway,    Isle     of     Man,    Street    Railway 

Journal 23:  357  March  5,  1904 

Milwaukee  Gas  Light  Company,  Engineering  News.     55:  28  Jan.  11,  1906 
Moulineaux  Power  Station,  Paris,  American  Elec- 
trician      16: 325  July,  1904 

Power    Station    of    the    Metropolitan    Railway, 

Engineer  (United  States) 37:  271  Nov.  15,  1900 

American  Electrician 12:112  March,  1900 

Monterey  Gas   and   Electric   Company,   Engineer 

(United  States) 42:87  Jan.  16,  1905 

Market  Street  Railway  Company  of  San  Francisco, 

Power 19:3  March,  1899 

Manhattan  Elevated  Railway,  Power 21: 1  April,  1901 

Street  Railway  Review 9:  82  Feb.  16,  1899 

Engineering  (United  States) 39:80  Feb.  1,  1902 

Engineering  News 47:  82  Jan.  30,  1902 

Milford,     Attleboro    and    Woonsocket    Railway, 

Street  Railway  Review 10:  638 


806 


STEAM  POWER  PLANT  ENGINEERING 


STEAM  ENGINE  PLANTS — Continued. 

New  York  Central  Power  House,  Electrical  World ...  48 :  95              July,  1905 

Power 24:228            April,  1904 

Newcastle  and  District  Electric  Lighting  Company, 

Engineer  (United  States) 40:  29              Jan.  1,  1903 

Northwestern  Elevated  Railroad,   Street  Railway 

Review 15:  199 

267 

17:655 
19:500 
23: 162 

Newcastle,  England,  Engineering  Record 46:  157            Aug.  16,  1902 

New   York  Gas  and  Electric   Light   and   Power 

Company,   Engineering  News 45 :  375            May  23,  1901 

Oakland,  California,  Engineering  Record 49 :  591             May  7,  1904 

Omaha    and  Council  Bluffs  Railway  and  Bridge 

Company,  Street  Railway  Review 9 :  298            May  15,  1899 

Paris  Metropolitan,  American  Electrician 16:  111             March,  1904 

Engineer  (United  States) 40:  13              Jan.  1,  1903 

Pan-American      Exposition,      Engineer      (United 

States) 38:158            June  1,  1901 

Paris  Exposition  Plant,  Power 19:  7                Dec.,  1899 

Pittsburg     Traction     Company,     Street     Railway 

Review 9:  135            Feb.  15,  1899 

15:  401  July  15,  1905 

Philadelphia     Rapid     Transit     Company,     Street 

Railway  Review 15:  526            Sept.  15,  1905 

Engineering  Record 47:  611            June  20,  1903 

52:340  Sept.  23,  1905 

Electrical     Power     Development     at     Portland, 

Oregon,  Electrical  World 48:  174            July  29,  1905 

Engineering  Record 52:  142            Aug.  5,  1905 

Peekskill  (New  York)  Light  and  Railway  Company, 

Street  Railway  Journal 20:  92              July  19,  1902 

Pacific  Electric  Railway  Company,  Los  Angeles, 

Street  Railway  Journal 23:  394            March  12,  1904 

Richmond,  Virginia,  Engineering  Record 49:  11              Jan.  2,  1904 

Rock    Island    Railway    Shops   at    East    Moline, 

Illinois,  Engineering  Record 50:  137            July  30,  1904 

Rochester  Power  Plant,  Engineer  (United  States).  38:  312             Sept.  1,  1901 
Scioto    Valley    Traction    Company,     Engineering 

Record 50:  644            Dec.  3,  1904 

St.    Louis    Exposition    Power   Plant,    Engineering 

News 42:  320 

American  Electrician 16:  528            Oct.,  1904 

Engineering  News 42:  223            Sept.  15,  1904 

Western  Electrician 35:  303            Oct.  15,  1904 

Power 23:  481            Nov.,  1903 

Saginaw  and  Bay  City  Light  Company,  Engineer 

(United  States) 42:  399            June  15,  1905 


APPENDIX  A 


807 


STEAM  ENGINE  PLANTS — Continued. 

St.     Louis     Transit     Company,     Street    Railway 

Journal 

Engineering  News 


Stark  Electric  Railway  Company,  Street  Railway 
Journal 

The  Sunderland  District  Tramways,  Street  Railway 
Journal 

St.  Clair  Tunnel,  Grand  Trunk  Railway,  Engineer- 
ing News 

Dayton,  Springfield  and  Urbana  Railway,  American 
Electrician 

Southern  Pacific  Company,  Galveston,  Texas, 
American  Electrician 

Syracuse  Power  Station,  Street  Railway  Journal .... 

Sydney,  New  South  Wales,  Street  Railway  Journal . . 

Seattle  Power  Station,  Street  Railway  Journal 

South  Side  Elevated,  Western  Electrician 

Springfield,  Illinois,  Light  and  Power  Company, 
Electrical  World 

Toledo  and  Monroe,  Street  Railway  Journal 

Tokio  Electric  Railway,  Street  Railway  Journal .  .  . 

Toledo,  Port  Clinton  and  Lakeside  Railway, 
Engineer  (United  States) 

Toledo  and  Western  Railway,  Street  Railway 
Journal 

Twin  City  Rapid  Transit  Company,  Engineering 

Record 

Street  Railway  Review 

Terminal  Railroad  Association  of  St.  Louis, 
Engineering  Record 

Trans-Mississippi  Exposition,  Omaha,  Power 

Toledo  and  Maumee  Valley  Railway,  Street  Rail- 
way Review 

American  Electrician 

United  Railways  and  Electric  Company,  Balti- 
more, Engineer  (United  States) 

Union  Traction  Company,  Street  Railway  Review .  . . 

Union  Traction  Company,  Anderson,  Indiana, 

Engineering  Record 

Street  Railway  Journal 

United  Electric  Company,  Hoboken,  New  Jersey, 
Engineering  Record 

Union  Loop  Power  House,  Chicago,  Western  Elec- 
trician   

Virginia  Electric  Railway  and  Equipment  Com- 
pany, Engineering  News 

Voltellina  Three-Phase  Railway,  Street  Railway 
Journal . . 


11:813 

47:  269 

297 

25:10 
26:96 
55:62 
12:415 

15:  476 
19:517 
20:930 
21:649 
26:231 

47:  253 
18:  124 
19:245 

43:295 
20:980 

50:692 
14:441 

51:92 

18:7 

9:613 
12:464 

39:357 
9:  12 

43:495 

18:  826 

46:56 

24:1 

45:318 

21:791 


Nov.  15,  1901 
April  3,  1902 
April  10,  1902 

Jan.  7,  1905 
July  15,  1905 
Jan.  18,  1906 
Sept.,  1900 

Oct.,  1903 
May,  1902 
Dec.  6,  1902 
May  2,  1903 
April  14,  1900 

Feb.  3,  1906 
Aug.,  1901 
March,  1902 

March,  1906 
Dec.  20,  1902 

Dec.  10,  1904 
July  20,  1904 

Jan.  28,  1905 
June,  1898 

Sept.  15,  1899 
Oct.,  1900 

May  15,  1902 
Jan.  15,  1899 

May  25,  1901 
Dec.,  1901 

July  19,  1902 
Jan.  7,  1899 
May  2,  1901 
May  30,  1903 


808 


STEAM  POWER   PLANT  ENGINEERING 


STEAM  ENGINE  PLANTS  —  Continued. 

Worcester  and  Blackstone  Valley  Street  Railway, 

Engineering  Record 47:  462  May  2,  1903 

Warren   and    Jamestown    Single-Phase    Railway, 

Street   Railway    Journal 27:  270  Feb.  16,  1906 

Electrical  World 47:  363  Feb.  17,  1906 

Washtenaw  Electric  Company,  Engineer  (United 

States) . . .     38: 150  May  15,  1901 

Willimantic  Gas  and  Electric  Company,  Engineer 

(United  States) 42:  163  March  1,  1905 

West  Albany  Power  Station,  Power 24:  677  Nov.,  1904 

Wilkesbarre   Interurban  Railway,   Street  Railway 

Review 9:  131  Feb.  15,  1899 

Weehauken    Station     New     York     Central     and 

Hudson  River  Railroad,  American  Electrician.  .      17:  501  Oct.,  1905 

Engineering  News 54:  506  Nov.  16,  1905 

Street  Railway  Journal 26:  872  Nov.  11,  1905 

Engineering  Record 52:  553  Nov.  11,  1905 

Western  Pennsylvania  Railway  and  Light  System, 

Street  Railway  Journal 20:  139  Aug.  2,  1902 

Zanesville  Railway  and  Light  Company,  Electrical 

World 43:500  March  19,  1904 

TURBINE  PLANTS. 

Amsterdam-Haarlem  Electric  Railway,  Engineer- 
ing Record 51:18  Jan.  17,  1905 

Street  Railway  Journal 25:  21  Jan.  7,  1905 

Brooklyn    Rapid    Transit,    Williamsburg    Plant, 

Street  Railway  Journal 26:  432  Sept.  23,  1905 

Generating  Station  Champ,  American  Electrician. .  17:  65  Feb.,  1905 
Clyde  Valley  Electric  Power  Company,  Engineering 

Record 52:  209  Sept.  9,  1905 

Chelsea  Generation  Station  of  the  London  Under- 
ground, Engineering  Record 52:  215  Feb.  25,  1905 

Power 23:421  Aug.,  1903 

Street  Railway  Journal 25;  388  March  4,  1905 

Commonwealth    Electric    Company,    Fisk    Street 

Station,   Western  Electrician 38:  55  Jan.  20,  1906 

Power 26:  715  Dec.,  1906 

Dubuque,  Iowa,  Power  Plant,  Engineering  Record.  20:  202  Aug.  13,  1904 

Detroit  Edison  Company,  Engineering  Record 52:  194  Oct.  7,  1905 

DeBeers  Consolidated  Mines,  Engineering  Record.  ..  51 :  4  Jan.  7,  1905 
Edison   Electric   Illuminating   Company,    Boston, 

Engineering  Record 51:150  Feb.  11,  1905 

Power 25:  389  July,  1905 

Hartford  Electric  Light  Company,  Dutch  Point 

Plant,  Engineering  Record 51:  204  Feb.  25,  1905 

Long  Island  Railroad  Power  House,  Street  Railway 

Journal 25:  24  Jan.  7,  1905 

Engineering  Record 49:  454  April  9,  1904 


APPENDIX  A  809 


TURBINE  PLANTS  —  Continued. 


Los  Angeles  Edison  Company,  Power 26:  67  Feb.,  1906 

Mexican  Central  Shops  at  Aguascalientes,  Engineer- 
ing Record 50:  227  Aug.  20,  1904 

247  Aug.  27,  1904 
Municipal    Turbine     Plant,     Anderson,     Indiana, 

Engineer  (United  States) 42:  641  Oct.  2,  1905 

Manchester,  England,  Power  Station,  Street  Rail- 
way Journal 25:  934  May  27,  1905 

New   York  Edison  Waterside   Station,    Electrical 

World 46:  383  .Sept.  2,  1905 

435  Sept.  9,  1905 

Power 22 :  1  Jan.,  1902 

New  York  Central  Steam-Electric  Station,  Power  26:  131  March,  1906 

Engineer  (United  States) 43:  733  Nov.  15,  1906 

New  York  and  Long  Island  River  Power  Station, 

Power 26:199  April,  1906 

New  Orleans  Power  House,  Power 24:  651  Nov.,  1904 

Street  Railway  Review 9:  393  June  15,  1899 

New    Bedford     Power    Station,     Street    Railway 

Review 11:  884  Dec.,  1901 

Old  Colony  Street  Railway,  Quincy  Point  Station, 

Engineering  Record 51 :  646  June  10,  1905 

Street  Railway  Journal 25:  1022  June  10,  1905 

Engineer  (United  States) 43:  85  Jan.  15,  1906 

Steam  Turbine  Plant  at  Poughkeepsie,  New  York, 

Shop,  Engineering  Record 51 :  454  April,  1905 

Potomac  Electric  Power  Company,   Washington, 

District  of  Columbia,  Power 27:  277  May,  1907 

Isolated  Stations. 

APARTMENTS. 

Collingwood      Apartment     Hotel,      New     York, 

Engineering  Record 45:  323  April  5,  1902 

Ansonia     Apartment,     New     York,     Engineering 

Record 46:  467  Nov.  15,  1902 

HOSPITALS. 

Agnes  Memorial  Sanatorium,  Engineering  Record ...     50:  312  Sept.  10,  1904 

Connecticut  Hospital  for  the  Insane,  Engineering 

Record 52:  44  July  8,  1905 

Lakeside  Hospital,  Cleveland,  Ohio,  Engineer 

(United  States) 39:  108  Feb.  15,  1902 

Massachusetts  General  Hospital,  Boston,  Transac- 
tions American  Society  of  Mechanical  Engineers  22:392  Jan.,  1901 

HOTELS. 

Bellevue-Stratford  Hotel,  Philadelphia,  Engineer- 
ing Record 51 :  14  Jan.  7,  1905 


810  STEAM  POWER  PLANT  ENGINEERING 

HOTELS  —  Continued. 

Hotel  Belmont,  New  York,  Engineering  Record 52 :  739  Dec.  30,  1905 

53:9  Jan.  6,  1906 

56  Jan.  13,  1906 

81  Jan.  20,  1906 

Hotel  Gotham,  New  York,  American  Electrician ....  17 :  551  Nov.,  1905 

Engineering  Record 52:  517  Nov.  4,  1905 

New  York  Athletic  Club,  Electrical  World 31:  463  April  16,  1898 

585  May  14,  1898 

University  Club,  Engineering  Record 46:  36  July  12,  1902 

MANUFACTURING  PLANTS. 

American     Steam     Pump     Company,     Engineer 

(United  States) 39:  217  April  1,  1902 

Atlas  Knitting  Mills,  Power 25:  408  July,  1905 

American      Lithographic      Company,      American 

Electrician 10 :  1 

Armour  Packing  Company  (J.  E.  Smith),  American 

Electrician 12 :  202  May,  1900 

Western  Electrician 23:  305  Nov.  26,  1898 

Anheuser-Busch    Brewing    Association,    Engineer 

(United  States) 42:  67  Feb.  1,  1905 

Power 25:84  Feb.,  1905 

Booth  Cold  Storage,  Engineering  Record 45:  97  Feb.  1,  1902 

Columbian  Cordage  Company,  Engineering  Record.  50:  447  Oct.  15,  1904 

Power  Plants  for  Cotton  Mills,  Engineer  (United 

States) 41:  135  Feb.  15,  1904 

Cascade  Water  Power  and  Light  Company,  Cas- 
cade, Canada,  Engineer  (United  States) 40:  603  Aug.  1,  1903 

Camden     Interstate     Railway     Company,     Street 

Railway  Review 15:269  May  15,  1905 

DeCew  Falls  Power  Plant,  Western  Electrician 38:  115  Feb.  10,  1906 

DeSabla,  California,  Station,  Engineering  News .  ...  54:  131  Aug.  10,  1905 

Dan  River  Power  and  Manufacturing  Company's 

Plant,  Engineering  Record 50:  291  Sept.  3,  1904 

Power  Plants  of  Edison  Electric  Company  of  Los 

Angeles,  Engineering  Record 51 :  211  Feb.  25,  1905 

302  March  11,  1905 

325  March  18,  1905 

Electric     Generating    Station   at    Glomen,     near 

Kykkelsrud,  American  Electrician 17:  409  Aug.,  1905 

Engineering  Record 50:  19  July  2,  1904 

Elgin  Watch  Works,  Elgin,  Illinois,  Engineering 

Record 56:  294  Sept.  14,  1907 

Hydro-Electric    Power   Development   for   Guana- 
juato, Mexico,  Engineering  Record 50:  195  Aug.  13,  1904 

General  Electric  Company's  Hudson  River  Plant, 

Electrical  World 43:  1115          June  11,  1904 

Garvins     Falls,     New     Hampshire,     Engineering 

Record 49:  668  May  28,  1904 


APPENDIX  A 


811 


MANUFACTURING  PLANTS  —  Continued. 


Gresivaudan   Valley  Power  Plant,    Western  Elec- 

trician   

Hudson  River  Company,  Mechanicsville,  New  York 

23:  168 

Oct.  1,  1898 

Engineering  News  

40:130 

Sept.  1,  1898 

Western  Electrician  

23:  135 

Sept.  3,  1898 

Hampton,  Virginia,  Engineering  Record  

48:  179 

Aug.  15,  1903 

Indian  Power  Plant,  Engineer  (London)  

101:36 

Jan.  12,  1906 

Engineering  

81:103 

Jan.  19,  1906 

Jhelum  Power  Plant,  Engineer  (United  States)  

43:240 

March  15,  1906 

Kern    River   Company's    Enterprise,    Engineering 

News  

52:55 

July  21,  1904 

Little  Falls,  Montana,  Power  Station,  Engineering 

Record  

51:616 

June  3,  1905 

McCormick  Twine  Mills,  Western  Electrician.  

28:  109 

Feb.  16,  1901 

National      Cash      Register     Company,      Dayton, 

Engineer  (United  States)  

39:  136 

March  1,  1902 

New  England   Confectionery  Company,   Engineer 

(United  States)  

41:231 

April  1,  1904 

Otis    Elevator    Company,    Yonkers,    New    York, 

Engineer  (United  States)  

39:327 

May  1,  1902 

Olympia  Cotton  Mills,  Engineer  (United  States)  .  .  . 

38:  382 

Oct.  15,  1901 

Power  Plant  of  a  Large  Silk  Mill,  Engineer  (United 

States)  

38:424 

Nov.  15,  1901 

Power  Plant  of  a  Sulphite  Mill,  Engineer  (United 

States)  

38:  445 

Dec.  1,  1901 

Swift  and  Company,  American  Electrician  

13:  149 

April,  1901 

Stickney   and   Poor  Spice   Factory,   Charlestown, 

Massachusetts,   American  Electrician  

14:207 

May,  1902 

United  Machinery  Company,  Engineering  Record  .  . 

52:  198 

Aug.  15,  1905 

Whittall  Mills,  Engineering  Record  

51:510 

May  6,  1905 

Washington    Mill    Company,    Lawrence,    Massa- 

chusetts, Engineering  

66:  533 

Oct.  21,  1898 

Western  Electric   Company,   New  York  Factory, 

American  Electrician  

12:333 

July,  1900 

Wet  more      Tobacco     Factory,     American     Elec- 

trician   

13:61 

Feb.,  1901 

Western    Wheel    Works,    Chicago,    Western   Elec- 

trician   

22:75 

Feb.  5,  1898 

Wood    Worsted    Mill,    Lawrence,    Massachusetts, 

Power  

27:73 

Feb.,  1907 

MISCELLANEOUS. 

Boston     South     Terminal     Station,     Engineering 

Record  

39:  346 

March  18,  1899 

Brooklyn  Institute  of  Arts  and  Science,  American 

Electrician  

17:  121 

March,  1905 

Birmingham  University  (C.  A.  Smith),  Engineering  . 

80:  341 

Sept.  15,  1905 

397 

Sept.  29,  1905 

507 

Oct.  20,  1905 

812 


STEAM  POWER  PLANT  ENGINEERING 


MISCELLANEOUS  —  Continued. 

Bryn   Mawr   College    (C.    G.    Gray),    Engineering 

Record 53:  183            Feb.  17,  1906 

University  of  California,  Engineer  (United  States) . .  42:  263            April  15,  1905 
Chicago,  Milwaukee  and  St.  Paul  Railway  Shops, 

Engineering  Record 48:  594            Nov.  14,  1903 

Columbia  University,  Engineering  Record 39:  546            May  13,  1899 

University  of  Chicago,  Engineering  Record 45:  246            March  15,  1902 

DuBois  Shops,  Engineering  Record 46:  218            Sept.  6,  1902 

Street  Railway  Journal 21 :  694            May  9,  1903 

Elizabethport  Railroad  Shops,  Engineering  Record . .  45:  581             June  21,  1902 

Hippodrome,  New  York,  Engineering  Record 52:  229            Aug.  26,  1905 

Harvard    Electric    Light    Plant    (W.    L.   Robb), 

American  Electrician 12:  107            March,  1900 

Lacka wanna  Railroad,  Engineering  Record 52:  507            Nov.  4,  1905 

Michigan   University    (H.    S.    Carhart),    Electrical 

World 31:  550            May  7,  1898 

Nassau,      Bahama      Islands,      Plant,       Electrical 

World 37:  914            June  1,  1901 

Ohio  State  University,  Engineer  (United  States)  .  .  39:  164            March  15,  1902 

Electrical  World 34:  1005          Dec.  30,  1899 

Princeton  University,  Engineer  (United  States) ...  41:  411            June  15,  1904 

Simmons  College,  Boston,  Engineering  Record 51:  161            Feb.  11,  1905 

Scranton  Schools,  Engineer  (United  States) 37:  148            June  1,  1900 

Power   Plant   of   a   University    (E.    A.    Darling), 

Transactions    American    Society    of  Mechanical 

Engineers 20:  663            1899 

Piping   Plans    for    the  Onondaga  County  Court- 

House,   Syracuse,  New  York,   Power 26:1                Jan.,  1906 

United  States  Bureau  of  Engraving,  Iron  Age 79 :  34              Jan.  3,  1907 

OFFICE  BUILDINGS. 

Arthur  Building,  New  York,  Engineering  Record.  .  50:  725            Dec.  17,  1904 
Atlantic  Building,   New   York,   Engineer   (United 

States) 38:262            July  1,  1901 

Arcade     Building,     Dayton,     Ohio,     Engineering 

Record 49:  767            June  18,  1904 

Broadway  Exchange  Building,  New  York,  Engineer 

(United  States) 38:  277            Aug.  1,  1901 

Engineering  Record 45:  374            Aug.  1,  1901 

Columbia  Office  Building,  Engineering  Record 48:  103 

Cable  Building,  Chicago,  Western  Electrician 25 :  377 

Chicago  and  Northwestern  Railway  Office  Building, 

Engineer  (United  States) 42:  739            Nov.  15,  1905 

Commercial    National    Bank    Building,    Chicago, 

Engineer  (United  States) 44                     Dec.  2,  1907 

Commerce  Realty  Building,  St.  Louis,  Engineering 

Record 39:  33              Dec.  10,  1898 

Ellicott    Square    Building,    St.    Louis,    Electrical 

World..                                  31:519 


APPENDIX  A  813 

OFFICE  BUILDINGS  —  Continued. 

Equipment  of  Tall  Office  Buildings,  in  New  York, 

Engineering  Record 39:  550 

Frick  Building,  Pittsburg,  Engineering  Record 45:  459 

Farmers'  Bank  Building,   Pittsburg,   Engineering 

Record 47:  492 

First  National  Bank  Building,  Chicago,  Engineer- 
ing Record 54:312  Sept.  22,  1906 

360  Sept.  29,  1906 

Flat-iron  Building,  New  York,  Engineer  (United 

States) 40:  296  April,  1903 

Federal    Building,     San    Francisco,     Engineering 

Record 47:  407,  578    April  18,  1903 

Heyworth  Building,  Engineer  (United  States) 42:  611  Sept.  15,  1905 

Isolated    Plant    for    Office    Buildings,    Electrical 

World .• 32:  108 

Isolated  Light  and  Power  Plant,   Western  Elec- 
trician      24: 910 

Consolidated  California  and  Virginia  Mining  Com- 
pany, American  Electrician 14:  425  Sept.,  1902 

Keystone    Bank    Building,    Pittsburg,    American 

Electrician 15: 171 

Kimball    Building,    Boston    (H.    S.    Knowlton), 

Engineer    (United   States) 42:  819  Dec.  15,  1905 

Large    and    Modern    Isolated     Plant,    American 

Electrician 15: 1 

Land    Title    and    Trust    Building,    Philadelphia, 

Electrical    World 32:  45 

Murphy  Power  Building,  Engineer  (United  States).     42:  67  Jan.  16,  1905 

Mutual  Life  Building,  Engineering  Record 47:  85 

Maiden  Lane  Building,   New   York,   Engineering 

Record 48:  770 

Methodist  Book  Concern,  Chicago,  Western  Elec- 
trician      22: 199 

Metropolitan  Life  Building,  New  York,  Engineer- 
ing Record 55:  97  Jan.  26,  1907 

Modern  Commercial  Building,  Electrical  World 32 :  623 

National    Bank    Building,     Pittsburg,     Engineer 

(United    States) 40:  387  June  1,  1903 

First  National  Bank,  Chicago,  Power 25:  297  May,  1905 

First  National   Bank,   Uniontown,   Pennsylvania, 

Engineering   Record 46: 13 

New  Building,  Dallas,  American  Electrician 14:  11  Jan.,  1902 

Oliver  Building,  Boston,  Engineering  Record 50:  717  Dec.  17,  1904 

Power    Plants    of    Office    Buildings,    Engineering 

News 51:  537 

Engineering  Record 49 :  725 

Power 24:419  July,  1904 

Transactions  American  Society  of  Mechanical 

Engineers 20:  880  July,  1899 


814 


STEAM  POWER  PLANT  ENGINEERING 


OFFICE  BUILDINGS  —  Continued. 

Power  Plants  of  the  Tall  Office  Buildings  (J.  H. 

Wells),  Engineering 78:130  July  22,  1904 

Engineering  Review 14:  19  July,  1904 

Power  Plants  of  Tall  Office  Buildings  (Wells  and. 
Bolton),  Transactions  American  Society  of  Mechan- 
ical Engineers 25: 685  June,  1904 

Phipps  Building,  Pittsburg,  Engineering  Record 50:  343  Sept.  17,  1904 

Power  Building,  in  Providence,  Engineering  Record  51: 162  Feb.  11,  1905 

Park  Row  Building,  Power 22 :  1  Oct.,  1902 

Electrical  World 34:  5 

Prudential      Building,      Newark,      New     Jersey, 

Engineering  Record 46:  367 

Pittsburg  and   Lake   Erie  Terminal,    Engineering 

Record 46:98  Aug.  2,  1902 

152  Aug.  16,  1902 

Pennsylvania       Railroad       Station,       Pittsburg, 

Engineering  Record 46:  203  Aug.  30,  1902 

Rose  Building,  Cleveland,  Engineer  (United  States)  38:  404  Nov.  1,  1901 

Railway    Exchange    Building,    Chicago,    Engineer 

(United  States) 41:  763  Nov.  15,  1904 

Rogers-Peet    Building,    New    York,    Engineering 

Record 46:  36  July  12,  1902 

Republic  Building,  St.  Louis,  American  Elec- 
trician   12:67  Feb.,  1900 

Rock  Island  Station,  Chicago,  Engineering  Record .  .  48:  328  Sept.  19,  1903 

Tribune  Building,  Chicago,  Engineering  Record ....  45:  607 

Western     Reserve     Building,     Engineer     (United 

States) 38:  87  March  1,  1901 

Wells    Building,    Milwaukee,     Engineer     (United 

States) 40:189  March  2,  1903 

STORES. 

Boston  Post  Office,  Electrical  World 46:  486  Sept.  16,  1905 

Boston  Store,  Chicago,  Engineer  (United  States) ...  44:  559  June  15,  1907 

Daniels  and  Fisher,  Denver,  Engineering  Record ...  50:  294  Sept.  3,  1904 

Ferguson-McKinney    Dry    Goods    Company,    St. 

Louis,    Engineer    (United   States) 40:  131  Feb.  2,  1903 

Marshall  Field  and  Company,  Engineering  Record ...  48 :  366  Sept.  26,  1903 

Western  Electrician 31 :  165  Sept.  13,  1902 

New  York  Federal  Building,  Electrical  World 31 :  379  March  25,  1896 

Government  Printing  Office,  Electrical  World 31:  94  Jan.  15,  1898 

115  Jan.  22,  1898 

Engineering  Record 47:  512  May  16,  1903 

543  May  23,  1903 

Modem  Printing  Plant,  Engineer  (United  States) ...  41 :  443  July  1 ,  1 904 

Toetz  Store,  Munich,  Engineering  Record 53:  160  Feb.  10,  1906 

Wanamaker  Store,  Engineering  Record 52:  94  July  22,  1905 

.  125  July  29,  1905 

53:219  Feb.  24,  1906 


APPENDIX  A 


815 


Power  Plant  Design. 

Power    Equipment    for  City  Roads,  Street  Railway 

Journal.. 22:  358  Aug.  29,  1903 

Production  and  Distribution  of  Alternating  Current 

for  Large  City  Systems,  Street  Railway  Journal ....     22:  506  Sept.  12,  1903 

Discussion 483 

Oil  Problem  in  Power  Stations,  Street  Railway  Journal     22:  664  Oct.  3,  1903 

Hydraulics    in    Connection    with    Street    Railway 

Operation,   Street  Railway   Journal 22:  709  Oct.  10,  1903 

Electric  Power  Generating  and  Transmission  Station, 

Street  Railway  Journal 24:  800  Oct.  29,  1904 

Wire  Glass  for  Power  Stations,  Street  Railway  Journal     24:  1049          Dec.  10,  1904 
Stark    Electric    Railway  Company,  Street    Railway 

Journal 25: 10  Jan.  7,  1905 

Electrification  of  the  London  Underground  Electric 
Railway  Company's  System  (S.  B.  Fortenbaugh), 

Street  Railway  Journal 25:  388  March  4,  1905 

Practical   Operation   and  Maintenance  of   Electrical 

Equipment,  Street  Railway  Journal 8:  33  Jan.,  1898 

Importance    of    the  Power    House,    Street    Railway 

Review 9:  458  July  15,  1899 

Transmission  of  Power,  Street  Railway  Journal 11:  24  Jan.  15,  1901 

11:154  March  15,  1901 

11:203  April  15,  1901 

11:363  June  15,  1901 

11:427  July  15,  1901 

11:497  Aug.  15,  1901 

11:669  Oct.  10,  1901 

11:831  Nov.  10,  1901 

Communication    in    Electric    Generating    Stations, 

Street  Railway  Journal 11 :  910  Dec.  15,  1901 

Power  House  Helps,  Street  Railway  Journal 12:  352  June  20,  1902 

Power  Speed  and  Efficiency  Curves,  Street  Railway 

Review 12:  534  Sept.  20,  1902 

Power    Transmission    and     Distribution    in     Utah, 

Street  Railway  Review 13:  288  May  20,  1903 

Power  for  Interurban  Lines,  Street  Railway  Review . .     13:  808  Oct.  20,  1903 

Power  Plant  Experiences,  Street  Railway  Review. ...      14:  898  Nov.  20,  1904 

Hints    in  Laying  Out    a     Power  Plant,   American 

Electrician , 13:  495  Oct.,  1901 

Power  Plant    of    a   Modern   Telephone    Exchange, 

American  Electrician 17:  190  April,  1905 

The    Power    Station    (F.   N.   Bushnell),    American 

Electrician 17:  561  Nov.,  1905 

The    Design    of    a    1,500-Kilowatt    Steam    Electric 
Central    Light    and    Power    Plant   (F.  Koester), 

Electrical  Review  (New  York) 50:  634  April  20,  1907 

670  April  27,  1907 

706  May  4,  1907 


816 


STEAM   POWER  PLANT  ENGINEERING 


Suggestions     for    Improvement     in     Power    Plants 

(A.  Bement),  Cassier's 13:  538  April,  1898 

Tendencies    in    Power   Plant   Design  (G.  L.  Clark), 

Cassier's .  .  . . 29:  433  March,  1906 

Standardizing  the  Power  Plant,  Electrical  Age 32 :  317  May,  1904 

Unipolar    Dynamos    and    Modern    Central   Station 

Design,  Electrical  Age 34:  114  Feb.,  1905 

Making  the  Most  of  an  Old  Plant,  Electrical  Age 34:  211  March,  1905 

Duplication     of     Electrical     Apparatus     in     Power 

Houses,  Electrical  Age 34:  443  June,  1905 

Design  of  Power  Stations,  Electrical  World 43:  991  May  21,  1904 

1030          May  28,  1904 
1090          June  4,  1904 
Design   of   an  Isolated   Power  and   Lighting  Plant, 

Electrical  World 47:  372  Feb.  17,  1906 

Piping  a  Steam  Plant,  Engineer  (United  States) 36:  78  April  1,  1899 

117  May  15,  1899 

Power      Plant     Specifications,     Engineer      (United 

States) 38:  293  Aug.  15,  1901 

Injudicious    Design     and    Management     of    Power 

Plants,  Engineer  (United  States) 40:  364  May  15,  1903 

Reversal     of    a     Large     Power     Station,    Engineer 

(United  States) 40:  876  Dec.  1,  1903 

Power   Plants    for   Cotton   Mills,  Engineer    (United 

States) 41 :  135  Feb.,  1904 

Power  Station  Design,  Engineering  Magazine 19:  902  Sept.,  1900 

Power    Plant    of    Apartment    Houses,    Engineering 

Magazine    26: 713  Feb.,  1904 

Power     Station     Design     (Merz     and     McLellan), 

Engineering    Magazine 27:  822  Aug.,  1904 

Modern     Power    Plant     Design     (Franz     Koester), 

Engineering  Magazine 29:  689  Aug.,  1905 

811  Sept.,  1905 

30:  71  Oct.,  1905 

182  Nov.,  1905 

Planning    and    Construction    of    the    Power    Plant 

(A.  E.  Dixon),  Engineering  Magazine  (Serial) 31 :  722  Aug.,  1906 

32:860  May,  1907 

Design  and  Construction  of  Modern  Central   Light 

Station     Power     House     (H.     H.     Humphrey), 

Engineering  News 43:  35  Jan.  18,  1900 

Design     of     Piping     for     Electric     Power    House, 

Engineering  Record 39 :  54  Dec.  17,  1898 

Power  Plant  in  a  Factory,  Engineering  Record 39:  430  April  8,  1899 

Design     and     Construction     of     Central     Station, 

Engineering  Record 40:  651  Dec.  9,  1899 

Foundation   of   a   Large   Power   Plant,  Engineering 

Record 40:  681  Dec.  16,  1899 

Design    of    Steam    Power    Plants    (H.   C.   Meyer), 

Engineering  Record 41 :  597  June  23,  1900 


APPENDIX  A 


817 


Design   of  Steam  Power  Plant   Condensers  (H.   C. 

Meyer,  Jr.),  Engineering  Record 43:  55  Jan.  19,  1901 

Feed  Water  Heaters  and  Economizers  (H.  C.  Meyer, 

Jr.),  Engineering  Record . 43:  224  March  9,  1901 

Mechanical  Draft  and  Chimneys  (H.  C.  Meyer,  Jr.), 

Engineering  Record 43:  468  May  18,  1901 

Some  Departures   in  Design  in  a  Detroit  Central 

Power  Station,  Engineering  Record 51 :  356  March  25,  1905 

Cement  in  Central  Station  Design,  Engineering  Record     51 :  481  April  29,  1905 

Compactness  in  Power    Plant    Design,   Engineering 

Record 51:  560  May  20,  1905 

Power  Station  Design  (F.  N.  Bushnell),  Engineering 

Record 52:  460  Oct.  21,  1905 

Metamorphosis  of  an  Electric  Power  Station,  Power     18:  1  June,  1898 

Power  Plant  Location,  Power 24 :  646  Nov.,  1904 

24:  770  Dec.,  1904 

Power  Plant  Location  (H.  D.  Jackson),  Power 25:  435  July,  1905 

Design  of   a  100-Horse-Power  Steam  Plant   (J.   F. 

Hobart),  Power 27:421  July,  1907 

Present  Trend  of  Station  Practice,   Street  Railway 

Journal 19:  750  June  14,  1902 

Designing  of  Steam  Power  Plants  (W.  C.  Kerr),  Street 

Railway  Journal 20:  497  Oct.  4,  1902 

Safety   Devices  in  Central   and   Substations,  Street 

Railway  Journal 21:  672  May  2,  1903 

Economical  and    Safe    Limits    on    Size    of    Central 

Stations,  Street  Railway  Journal 27:  671  May  2,  1903 

Design  of  City  Power  Stations,  Street  Railway  Journal    22:  282  Aug.  29,  1903 

Evolution  of  Modern  Power  Station,  Street  Railway 

Journal 24:  555  Oct.  8,  1904 

Size  of  Power  Stations,  Street  Railway  Journal 26: 131  July  22,  1905 

Power  Station  (F.  N.  Bushnell),  Street  Railway  Journal  Sept.  30,  1905 

Underground  Distribution  of  Power  for  Urban  Electric 

Station  (James  Heywood),  Street  Railway  Journal. .  .26:  269  Aug.  19  1905 

Designing  Boilers  for  a  Small  Street  Railway  Plant, 

Street  Railway  Review 9: 125  Feb.,  1899 

9:  188  March,  1899 

9:263  April,  1899 

Storage  Batteries  and  Railway  Power  Stations,  Street 

Railway  Review 9: 190  1899 

Bibliography,  1908-1910. 

GAS  ENGINE  PLANTS. 

Bituminous  Power  Producer  Plant: 

Pro.  Engr.  Soc.  W.  Penn 25:  603  Jan.,  1910 

Engineering,  London 108:  317  Sept.  24,  1909 

Engineering  Record 59:  817  June  26,  1909 

Engineering  News 61 :  42  Jan.  14,  1909 

Jour.  A.S.M.E 31:  1343          Dec.,  1909 

Blast  Furnace  Power  Plant,  Engineering  and  Mining 

Journal 87:  20  Jan.  2,  1909 


818 


STEAM  POWER  PLANT  ENGINEERING 


GAS  ENGINE  PLANTS—  Continued. 

Boston  Elevated  R.  R.  Co.'s  Plant,  Engineering 

Record 59:  352  March  27,  1909 

Carnegie  Steel  Plant,  Power  and  Engineer 28:  639  April  28,  1908 

Central    Station    at    Harvard,    111.,    Electrical 

World 53:  913  April  15,  1909 

Central   Station   at   Aberdeen,  South  Dakota, 

.      Electrical  World 54:  1567  Dee.  30,  1909 

Combined  Steam  and  Gas  Power  Plant:  Engi- 
neering Record 60:  329  Sept.  18,  1909 

Costs  of  a  Gas  Engine  and  of  a  Combined  Steam 

Plant,  Engineering  Record 60:  272  Sept.  4,  1909 

Gas  Engine  in  Central  Station  Work,  Iron  Age       84:  256  July  22,  1909 

Gas-Electric  Plant  at  Gary,  Indiana,  Power  and 

Engineer 29:  268  Aug.  18,  1908 

Western  Electrician 43:  245  Oct.  3,  1908 

Gas-Driven   Blowing   Plant  at  Gary,  Indiana, 

Engineering  Record 59:  274  March  G,  1909 

Gas  Power  vs.  Steam  Power  for  Central  Stations, 

Electrical  Review 65:  764  Nov.  12,  1909 

Philadelphia  Gear  Works,  American  Machinery       32:  155  July  22,  1909 

Recent  Development  of  the  Producer  Gas  Power 

Plant  in  the  United  States,  U.  S.  Geological 

Survey Bui.  416  No.  11142 

Small  Isolated  Producer  Gas  Plant,  Power  and       32:  68  Jan.  1,  1910 

Engineer 31:  697  Oct.  26,  1909 

Working    Results,    Gas-Electric    Power  Plant, 

Pro.  Am.  Inst.  Eiec.  Engrs 27:  1123  July  1,  1909 

HYDRO-ELECTRIC. 

Hennepin  Power  House,  St.  Anthony  Falls, 

Engineering  Record 59:  676  May  29,  1909 

La  Crosse,  Wis.,  Water  Power  Company,  Elec- 
trical World 55:  801  March  31,  1910 

Lansing,  Mich.,  Hydro-Electric  with  Steam 

Auxiliaries,  Electrical  Review,  New  York 55:  1235  Dec.  25,  1909 

Low-Head  Hydro-Electric,  Tippecanoe  River, 

Monticello,  Ind.,  Electrical  World 54:  975  Oct.  21,  1909 

Rainbow  Falls,  Engineering  Record 61 :  292  March  12,  1910 

Sioux  Falls,  Electrical  World 53:  963  April  22,  1909 

Telluride  Power  Company,  Bear  River  Plant, 

Engineering  Record 61 :  353  March  26,  1910 

Wenitchee  Electric  Company,  Journal  of  Elec- 
tricity Power  and  Gas  23:  87  July  31,  1909 

STEAM  ENGINE  —  CENTRAL  STATIONS. 

Central  Pennsylvania  Traction  Company,  Elec- 
tric Railway  Review 19:  455  April  11,  1908 

Commonwealth  Power  Company,  Power  and 

Engineer 31:  127  July  27,  1909 


APPENDIX  A 


819 


STEAM  ENGINE  —  CENTRAL  STATIONS  —  Continued. 

Home-Electric  and  Steam  Heating  Company, 

Electrical  World 54:  1345  Dec.  2,  1909 

Indianapolis,  Crawfordsville  and  Western  Trac- 
tion System,  Street  Railway  Journal 31 :  850  May  23,  1908 

Louisville  Lighting  Company,  Power 30:  663  Apr.  13,  1909 

Redonda  Power  Plant,  Redonda,  Cal.,  Electrical 
Railway  Journal 32:  618  Sept.  12,  1908 

United  Railways  of  Baltimore,  Street  Railway 
Journal 31 :  770  May  9,  1908 

Williamsport,  Penn.,  New  Plant  at,  Power  and 

Engineering 31:  213  Aug.  10,  1909 

STEAM  TURBINE  PLANTS. 

Brockton,  New  Lighting  Station  at,  Power  and 

Engineer 30:  315  Feb.  16,  1909 

Commonwealth     Edison     Company,     Chicago, 

Fisk  Street,  Electrical  World 51 :  1023  May  16,  1908 

Northwest  Station,  Electrical  World 55:  667  March  17,  1909 

Quarry  Street,  Electrical  World 53:  17  Jan.  2,  1909 

Double  Flow  Turbines  at  Brunat  Island,  Street 

Railway  Journal 31 :  908  May  30,  1908 

Engineering  Record 57:  693  May  30,  1908 

Government  Plant  at  Washington,  Power  and 

Engineer 32:  838  May  10,  1910 

Hartford  Electric  Light  Company,  Power  and 

Engineer 31 :  884  Nov.  30,  1909 

Hoboken  Passenger  Terminal  of    the    Lacka- 

wanna  Railway,  Engineering  Record 59:  184  Feb.  13,  1909 

Jackson    Electric   Railway,  Light   and  Power 

Company,  Street  Railway  Journal 31 :  278  Feb.  22,  1908 

New  York  Edison,  Waterside  No.  2,  Electrical 

World 53:  545  March  4,  1909 

Knoxville  Railway  and  Light  Company,  Electric 

Railway  Journal 32:  534  Aug.  29,  1908 

Norfolk  Traction  Power  Plant,  Power  and  Engi- 
neer   32:  702  April  19,  1910 

Recent  Electric  Railway  Power  Station  Design, 

Electrical  Railway  Journal 33:  538  March  27,  1909 

Sayre  Electric  Company,  Electrical  World 55:  273  Feb.  3,  1910 

ISOLATED  STATIONS. 

Apartment  Buildings,  Apthorp  Apartment,  New 

York,  Engineering  Record 60:  69  July  17,  1909 

Belnord  Apartment,  New  York,  Domestic  Engi- 
neering    Dec.  25,  1909 

HOSPITALS. 

Oak  Park  Infirmary,  Electrical  World 53:  912  April  15,  1909 

St.  Luke's,  Chicago,  Power  and  Engineer 29:  996  Dec.  15,  1908 


820 


HOTELS. 


STEAM  POWER   PLANT  ENGINEERING 


La  Salle,  Chicago,  Electrical  Review  and  West- 
ern Electrician ..'... 55:  418  Sept.  4,  1909 

Engineering  Record 60:  241  Aug.  20,  1909 

New  Plaza,  New  York,  Engineering  Record 58:  577  Nov.  21,  1908 

MISCELLANEOUS. 

Carnegie  Institute,  Pittsburg,  Power  and  Engi- 
neer   30:  97  Jan.  12,  1909 

West  Point  Military  Academy,  Power  and  Engi- 
neer   . . . . . 30:  747  April  27,  1909 

MANUFACTURING  PLANTS. 

Allis-Chalmers  Company,  Milwaukee,  Engineer- 
ing Record 61 :  194  Feb.  12,  1910 

Corn  Products  Company,  Argo,  111.,  Practical 

Engineer  (United  States) 14:336  June,  1910 

Coronet     Phosphate     Company,     Engineering 

Record . 60:  469  Oct.  23,  1909 

Johnson  &   Johnson,  New  Brunswick,   N.   J., 

Power  and  Engineer..  . 32:  568  March  29,  1910 

Heath  &   Milligan   Paint  Works,   Engineering 

Record , 59:  202  Feb.  20,  1909 

Sawmills,  Power  Plants  for,  Power 28:  973  June  23,  1908 

Textile  Mills  Power  Plants,  Electrical  Review .  .  .       53:  638  ;  *&      Oct.  31,  1908 

Jan.  2,  1909 

Cassier's  Magazine •>..,..„ ,..,;..       34:  371  Aug.,  1908 

Engineering  Record.... 58:  440  Oct.  17,  1908 

OFFICE  BUILDINGS. 

City  Investing  Building,  New  York,  Power  and 

Engineer ,  .  .  .       29:  983  Dec.  15,  1908 

Commercial  National  Bank,  Chicago,  Western 

Electrician 41:  441  Dec.  7,  1907 

Kuper      Building,      Baltimore,       Engineering 

Record 60:  492  Oct.  30,  1909 

Union    National    Bank    Building,     Pittsburg, 

Engineering  Record 57:  818  June  27,  1908 

58:  13  July  4,  1908 

POWER  PLANT  DESIGN. 

Central  Electric  Plants  in  Small  Towns,  Engi- 
neering News 62:  32  Sept.  23,  1909 

Central  Stations  vs.  Private  Plants,  Engineering       87:  288  Feb.  26,  1909 

Central  Stations,  Towns  over  1000  Inhabitants, 

Electrical  Review,  New  York 56:  83  Jan.  8,  1910 

Design  of  Steam  Electric  Plants  (Frank  Koes- 

ter),  Electrical  Review  (Serial) 54:  472  March  13,  190? 

From  an  Insurance  Standpoint,  Engineering 

Review 59:  295  March  13,  1909 

Heat  Losses  in  an  Electric  Power  Station, 

Engineering 87:  10  Jan.  1,  1909 


APPENDIX  A  821 

POWER  PLANT  DESIGN  —  Continued. 

Modern  Power  Station  Design  (H.  B.  Parsons), 

Cassier's  Magazine 36:  328  Aug.,  1909 

Prime    Movers    (C.    P.  Steinmetz),    Pro.  Am. 

Inst.  Elec.  Engrs :. 28:  135  Feb.,  1909 

Progress  in  Power  Plant  Auxiliary  Equipment, 

Street  Railway  Journal 31 :  36  Jan.  11,  1908 

Simplicity  in  Steam  Plant  Design,  Power 29:  371  Sept.  1,  1908 


APPENDIX  B. 

RULES  FOR  CONDUCTING  BOILER  TRIALS.* 
Code  of  1899. 

I.  Determine  at  the  outset  the  specific  object  of  the  proposed  trial, 
whether  it  be  to  ascertain  the  capacity  of  the  boiler,  its  efficiency  as  a 
steam   generator,   its  efficiency  and  its   defects   under  usual   working 
conditions,  the  economy  of  some  particular  kind  of  fuel,  or  the  effect  of 
changes  of  design,  proportion,  or  operation;  and  prepare  for  the  trial 
accordingly.     (Appendix  II.) 

II.  Examine  the  boiler,  both  outside  and  inside;  ascertain  the  dimen- 
sions of  grates,  heating  surfaces,  and  all  important  parts;  and  make  a 
full  record,  describing  the  same,  and  illustrating  special  features  by 
sketches.     The  area  of  heating  surface  is  to  be  computed  from  the  sur- 
faces of  shells,  tubes,  furnaces,  and  fire  boxes  in  contact  with  the  fire  or 
hot  gases.     The  outside  diameter  of  water  tubes  and  the  inside  diameter 
of  fire  tubes  are  to  be  used  in  the  computation.     All  surfaces  below  the 
mean  water  level  which  have  water  on  one  side  and  products  of  com- 
bustion on  the  other  are  to  be  considered  as  water-heating  surface,  and 
all  surfaces  above  the  mean  water  level  which  have  steam  on  one  side 
and  products  of  combustion  on  the  other  are  to  be  considered  as  super- 
heating surface. 

III.  Notice  the  general  condition  of  the  boiler  and  its  equipment,  and 
record  such  facts  in  relation  thereto  as  bear  upon  the  objects  in  view. 

If  the  object  of  the  trial  is  to  ascertain  the  maximum  economy  or 
capacity  of  the  boiler  as  a  steam  generator,  the  boiler  and  all  its  appur- 
tenances should  be  put  in  first-class  condition.  Clean  the  heating  surface 
inside  and  outside,  remove  clinkers  from  the  grates  and  from  the  sides 
of  the  furnace.  Remove  all  dust,  soot,  and  ashes  from  the  chambers, 
smoke  connections,  and  flues.  Close  air  leaks  in  the  masonry  and  poorly 
fitted  cleaning  doors.  See  that  the  damper  will  open  wide  and  close 
tight.  Test  for  air  leaks  by  firing  a  few  shovels  of  smoky  fuel  and 

*  From  the  report  of  the  Committee  of  the  American  Society  of  Mechanical 
Engineers  on  the  revision  of  the  Society  Code  of  1885  relative  to  a  standard  method 
of  conducting  steam  boiler  trials. 

822 


APPENDIX  B  823 

immediately  closing  the  damper,  observing  the  escape  of  smoke  through 
the  crevices,  or  by  passing  the  flame  of  a  candle  over  cracks  in  the 
brickwork. 

IV.  Determine  the  character  of  the  coal  to  be  used.     For  tests  of  the 
efficiency  or  capacity  of  the  boiler  for  comparison  with  other  boilers  the 
coal  should,  if  possible,  be  of  some  kind  which  is  commercially  regarded 
as  a  standard.     For  New  England  and  that  portion  of  the  country  east 
of  the  Allegheny  Mountains,  good  anthracite  egg  coal,  containing  not 
over  10  per  cent  of  ash,  and  semi-bituminous  Clearfield  (Pa.),  Cumber- 
land (Md.),  and  Pocahontas  (Va.)  coals  are  thus  regarded.     West  of  the 
Allegheny   Mountains,    Pocahontas    (Va.)    and   New   River   (W.    Va.) 
semi-bituminous,  and  Youghiogheny  or  Pittsburg  bituminous  coals  are 
recognized  as  standards.*     There  is  no  special  grade  of  coal  mined  in 
the  Western  States  which  is  widely  recognized  as  of  superior  quality 
or  considered  as  a  standard  coal  for  boiler  testing.     Big  Muddy  lump,  an 
Illinois  coal  mined  in  Jackson  County,  111.,  is  suggested  as  being  of 
sufficiently  high  grade  to  answer  the  requirements  in  districts  where 
it  is  more  conveniently  obtainable  than  the  other  coals   mentioned 
above. 

For  tests  made  to  determine  the  performance  of  a  boiler  with  a  par- 
ticular kind  of  coal,  such  as  may  be  specified  in  a  contract  for  the  sale 
of  a  boiler,  the  coal  used  should  not  be  higher  in  ash  and  in  moisture 
than  that  specified,  since  increase  in  ash  and  moisture  above  a  stated 
amount  is  apt  to  cause  a  falling  off  of  both  capacity  and  economy  in 
greater  proportion  than  the  proportion  of  such  increase. 

V.  Establish  the  correctness  of  all  apparatus  used  in  the  test  for  weigh- 
ing and  measuring.     These  are: 

1.  Scales  for  weighing  coal,  ashes,  and  water. 

2.  Tanks,  or  water  meters  for  measuring  water.     Water  meters,  as 
a  rule,  should  only  be  used  as  a  check  on  other  measurements.     For 
accurate  work,  the  water  should  be  weighed  or  measured  in  a  tank. 
(Appendices  I,  IV,  VII,  VIII.) 

3.  Thermometers  and  pyrometers  for  taking  temperatures  of  air, 
steam,  feed  water,  waste  gases,  etc.     (Appendix  XXVII.) 

4.  Pressure  gauges,  draught  gauges,  etc.    (Appendices  XXVIII  to 
XXX.) 

*  These  coals  are  selected  because  they  are  about  the  only  coals  which  contain 
the  essentials  of  excellence  of  quality,  adaptability  to  various  kinds  of  furnaces, 
grates,  boilers,  and  methods  of  firing,  and  wide  distribution  and  general  accessibility 
in  the  markets. 


824  STEAM  POWER  PLANT  ENGINEERING 

The  kind  and  location  of  the  various  pieces  of  testing  apparatus  must 
be  left  to  the  judgment  of  the  person  conducting  the  test;  always  keeping 
in  mind  the  main  object,  i.e.,  to  obtain  authentic  data. 

VI.  See  that  the  boiler  is  thoroughly  heated  before  the  trial  to  its  usual 
working  temperature.     If  the  boiler  is  new  and  of  a  form  provided  with 
a  brick  setting,  it  should  be  in  regular  use  at  least  a  week  before  the 
trial,  so  as  to  dry  and  heat  the  walls.     If  it  has  been  laid  off  and  become 
cold,  it  should  be  worked  before  the  trial  until  the  walls  are  well  heated. 

VII.  The  boiler  and  connections  should  be  proved  to  be  free  from 
leaks  before  beginning  a  test,  and  all  water  connections,  including  blow 
and  extra  feed  pipes,   should  be   disconnected,   stopped  with  blank 
flanges,  or  bled  through  special  openings  beyond  the  valves,  except  the 
particular  pipe  through  which  water  is  to  be  fed  to  the  boiler  during  the 
trial.     During  the  test  the  blow-off  and  feed  pipes  should  remain  exposed 
to  view. 

If  an  injector  is  used,  it  should  receive  steam  directly  through  a  felted 
pipe  from  the  boiler  being  tested.* 

If  the  water  is  metered  after  it  passes  the  injector,  its  temperature 
should  be  taken  at  the  point  where  it  leaves  the  injector.  If  the  quan- 
tity is  determined  before  it  goes  to  the  injector  the  temperature  should 
be  determined  on  the  suction  side  of  the  injector,  and  if  no  change  of 
temperature  occurs*  other  than  that  due  to  the  injector,  the  tempera- 
ture thus  determined  is  properly  that  of  the  feed  water.  When  the 
temperature  changes  between  the  injector  and  the  boiler,  as  by  the  use 
of  a  heater  or  by  radiation,  the  temperature  at  which  the  water  enters  and 
leaves  the  injector  and  that  at  which  it  enters  the  boiler  should  all  be 
taken.  In  that  case  the  weight  to  be  used  is  that  of  the  water  leaving 
the  injector,  computed  from  the  heat  units  if  not  directly  measured, 
and  the  temperature  that  of  the  water  entering  the  boiler. 

Let  w  =  weight  of  water  entering  the  injector. 
x    =  weight  of  steam  entering  the  injector. 
&4  =  heat  units  per  pound  of  water  entering  injector. 
h2  =  heat  units  per  pound  of  steam  entering  injector. 
h3  =  heat  units  per  pound  of  water  leaving  injector. 

*  In  feeding  a  boiler  undergoing  test  with  an  injector  taking  steam  from  another 
boiler,  or  from  the  main  steam  pipe  from  several  boilers,  the  evaporative  results 
may  be  modified  by  a  difference  in  the  quality  of  the  steam  from  such  source  com- 
pared with  that  supplied  by  the  boiler  being  tested,  and  in  some  cases  the  connection 
to  the  injector  may  act  as  a  drip  for  the  main  steam  pipe.  If  it  is  known  that  the 
steam  from  the  main  pipe  is  of  the  same  pressure  and  quality  as  that  furnished  by 
the  boiler  undergoing  the  test,  the  steam  may  be  taken  from  such  main  pipe. 


APPENDIX  B  825 

Then  w  +  x  —  weight  of  water  leaving  injector. 


See  that  the  steam  main  is  so  arranged  that  water  of  condensation 
cannot  run  back  into  the  boiler. 

VIII.  Duration  of  the  Test.  —  For  tests  made  to  ascertain  either  the 
maximum  economy  or  the  maximum  capacity  of  a  boiler,  irrespective 
of  the  particular  class  of  service  for  which  it  is  regularly  used,  the 
duration  should  be  at  least  10  hours  of  continuous  running.     If  the 
rate  of  combustion  exceeds  25  pounds  of  coal  per  square  foot  of  grate 
surface  per  hour,  it  may  be  stopped  when  a  total  of  250  pounds  of  coal 
has  been  burned  per  square  foot  of  grate. 

In  cases  where  the  service  requires  continuous  running  for  the 
whole  24  hours  of  the  day,  with  shifts  of  firemen  a  number  of  times 
during  that  period,  it  is  well  to  continue  the  test  for  at  least  24 
hours. 

When  it  is  desired  to  ascertain  the  performance  under  the  working 
conditions  of  practical  running,  whether  the  boiler  be  regularly  in  use 
24  hours  a  day  or  only  a  certain  number  of  hours  out  of  each  24,  the 
fires  being  banked  the  balance  of  the  time,  the  duration  should  not  be 
less  than  24  hours. 

IX.  Starting  and  Stopping  a   Test.  —  The  conditions  of  the  boiler 
and  furnace  in  all  respects  should  be,  as  nearly  as  possible,  the  same  at 
the  end  as  at  the  beginning  of  the  test.     The  steam  pressure  should  be 
the  same;  the  water  level  the  same;  the  fire  upon  the  grates  should 
be  the  same  in  quantity  and  condition;    and    the  walls,   flues,  etc., 
should  be  of  the  same  temperature.     Two  methods  of  obtaining  the 
desired  equality  of  conditions  of  the  fire   may   be  used,  viz.:    those 
which  were  called  in  the  Code  of  1885  "  the  standard  method  "  and 
"  the  alternate  method,"  the  latter  being  employed  where  it  is  incon- 
venient to  make  use  of  the  standard  method.* 

X.  Standard  Method  of  Starting  and  Stopping  a  Test.  —  Steam  being 
raised  to  the  working  pressure,  remove  rapidly  all  the  fire  from  the 
grate,  close  the  damper,  clean  the  ash  pit,  and  as  quickly  as  possible 
start  a  new  fire  with  weighed  wood  and  coal,  noting  the  time  and  the 

*  The  Committee  concludes  that  it  is  best  to  retain  the  designations  "standard" 
and  "alternate,"  since  they  have  become  widely  known  and  established  in  the 
minds  of  engineers  and  in  the  reprints  of  the  Code  of  1885.  Many  engineers  prefer 
the  "alternate"  to  the  "standard"  method  on  account  of  its  being  less  liable  to 
error  due  to  cooling  of  the  boiler  at  the  beginning  and  end  of  a  test. 


826  STEAM  POWER   PLANT  ENGINEERING 

water  level*  while  the  water  is  in  a  quiescent  state,  just  before  lighting 
the  fire. 

At  the  end  of  the  test  remove  the  whole  fire,  which  has  been  burned 
low,  clean  the  grates  and  ash  pit,  and  note  the  water  level  when  the 
water  is  in  a  quiescent  state,  and  record  the  time  of  hauling  the  fire. 
The  water  level  should  be  as  nearly  as  possible  the  same  as  at  the 
beginning  of  the  test.  If  it  is  not  the  same,  a  correction  should  be 
made  by  computation,  and  not  by  operating  the  pump  after  the  test  is 
completed. 

XI.  Alternate  Method  of  Starting  and  Stopping  a  Test.  —  The  boiler 
being  thoroughly  heated  by  a  preliminary  run,   the  fires   are  to  be 
burned  low  and  well  cleaned.     Note  the  amount    of   coal  left  on  the 
grate  as  nearly  as  it  can  be  estimated;  note  the  pressure  of  steam  and 
the  water  level.     Note  the  time,  and  record  it  as  the  starting  time. 
Fresh  coal  which  has  been  weighed  should  now  be  fired.     The  ash  pits 
should  be  thoroughly  cleaned  at  once  after  starting.     Before  the  end 
of  the  test  the  fires  should  be  burned  low,  just  as  before  the  start,  and 
the  fires  cleaned  in  such  a  manner  as  to  leave  a  bed  of  coal  on  the 
grates  of  the  same  depth,  and  in  the  same  condition,  as  at  the  start. 
When  this  stage  is  reached,  note  the  time  and  record  it  as  the  stopping 
time.     The   water   level   and   steam   pressures   should   previously   be 
brought  as  nearly  as  possible  to  the  same  point  as  at  the  start.     If  the 
water  level  is  not  the  same  as  at  the  start,  a  correction  should  be 
made  by  computation,  and  not  by  operating  the  pump  after  the  test 
is  completed. 

XII.  Uniformity   of  Conditions.  —  In  all  trials   made  to   ascertain 
maximum  economy  or  capacity,  the  conditions  should  be  maintained 
uniformly  constant.     Arrangements  should  be  made  to  dispose  of  the 
steam  so  that  the  rate  of  evaporation  may  be  kept  the  same  from 
beginning  to  end.     This   may  be  accomplished  in   a  single  boiler  by 
carrying  the  steam  through  a  waste  steam  pipe,  the  discharge  from 
which  can  be  regulated  as  desired.     In  a  battery  of  boilers,  in  which 
only  one  is   tested,   the   draft    can    be    regulated   on   the  remaining 
boilers,  leaving   the   test   boiler   to  work   under   a    constant   rate  of 
production. 

Uniformity  of  conditions  should  prevail  as  to  the  pressure  of  steam, 
the  height  of  water,  the  rate  of  evaporation,  the  thickness  of  fire,  the 

*  The  gauge  glass  should  not  be  blown  out  within  an  hour  before  the  water 
level  is  taken  at  the  beginning  and  end  of  a  test,  otherwise  an  error  in  the  reading 
of  the  water  level  may  be  caused  by  a  change  in  the  temperature  and  density  of  the 
water  in  the  pipe  leading  from  the  bottom  of  the  glass  into  the  boiler. 


APPENDIX  B  827 

times  of  firing  and  quantity  of  coal  fired  at  one  time,  and  as  to  the 
intervals  between  the  times  of  cleaning  the  fires. 

The  method  of  firing  to  be  carried  on  in  such  tests  should  be  dic- 
tated by  the  expert  or  person  in  responsible  charge  of  the  test,  and 
the  method  adopted  should  be  adhered  to  by  the  fireman  throughout 
the  test. 

XIII.  Keeping  the  Records.  —  Take  note  of  every  event  connected 
with  the  progress  of  the  trial,  however  unimportant  it  may  appear. 
Record  the  time  of  every  occurrence  and  the  time  of  taking  every 
weight  and  every  observation.     (Appendices  I,  IV,  V,  VI,  VII,  VIII.) 

The  coal  should  be  weighed  and  delivered  to  the  fireman  in  equal 
proportions,  each  sufficient  for  not  more  than  one  hour's  run,  and  a 
fresh  portion  should  not  be  delivered  until  the  previous  one  has  all 
been  fired.  The  time  required  to  consume  each  portion  should  be 
noted,  the  time  being  recorded  at  the  instant  of  firing  the  last  of  each 
portion.  It  is  desirable  that  at  the  same  time  the  amount  of  water 
fed  into  the  boiler  should  be  accurately  noted  and  recorded,  including 
the  height  of  the  water  in  the  boiler  and  the  average  pressure  of 
steam  and  temperature  of  feed  during  the  time.  By  thus  recording 
the  amount  of  water  evaporated  by  successive  portions  of  coal,  the 
test  may  be  divided  into  several  periods  if  desired,  and  the  degree 
of  uniformity  of  combustion,  evaporation,  and  economy  analyzed  for 
each  period.  In  addition  to  these  records  of  the  coal  and  the  feed 
water,  half-hourly  observations  should  be  made  of  the  temperature  of 
the  feed  water,  of  the  flue  gases,  of  the  external  air  in  the  boiler  room, 
of  the  temperature  of  the  furnace  when  a  furnace  pyrometer  is  used, 
also  of  the  pressure  of  steam,  and  of  the  readings  of  the  instruments 
for  determining  the  moisture  in  the  steam.  A  log  should  be  kept  on 
properly  prepared  blanks  containing  columns  for  record  of  the  various 
observations.  (Appendix  XXII.) 

When  the  "  standard  method  "  of  starting  and  stopping  the  test  is 
used,  the  hourly  rate  of  combustion  and  of  evaporation  and  the  horse 
power  should  be  computed  from  the  records  taken  during  the  time 
when  the  fires  are  in  active  condition.  This  time  is  somewhat  less 
than  the  actual  time  which  elapses  between  the  beginning  and  end  of 
the  run.  The  loss  of  time  due  to  kindling  the  fire  at  the  beginning  and 
burning  it  out  at  the  end  makes  this  course  necessary. 

XIV.  Quality  of  Steam.  —  The  percentage  of  moisture  in  the  steam 
should  be  determined  by  the  use  of  either  a  throttling  or  a  separating 
steam  calorimeter.     The  sampling  nozzle  should  be  placed  in  the  ver- 
tical steam  pipe  rising  from  the  boiler.     It  should  be  made  of  i-inch 


828  STEAM   POWER   PLANT   ENGINEERING 

pipe,  and  should  extend  across  the  diameter  of  the  steam  pipe  to 
within  half  an  inch  of  the  opposite  side,  being  closed  at  the  end  and 
perforated  with  not  less  than  twenty  J-inch  holes  equally  distributed 
along  and  around  its  cylindrical  surface,  but  none  of  these  holes  should 
be  nearer  than  J  inch  to  the  inner  side  of  the  steam  pipe.  The  calo- 
rimeter and  the  pipe  leading  to  it  should  be  well  covered  with  felting. 
Whenever  the  indications  of  the  throttling  or  separating  calorimeter 
show  that  the  percentage  of  moisture  is  irregular,  or  occasionally  in 
excess  of  three  per  cent,  the  results  should  be  checked  by  a  steam 
separator  placed  in  the  steam  pipe  as  close  to  the  boiler  as  convenient, 
with  a  calorimeter  in  the  steam  pipe  just  beyond  the  outlet  from 
the  separator.  The  drip  from  the  separator  should  be  caught 
and  weighed,  and  the  percentage  of  moisture  computed  therefrom 
added  to  that  shown  by  the  calorimeter.  (See  Appendices  XV  to 
XVII.) 

Superheating  should  be  determined  by  means  of  a  thermometer 
placed  in  a  mercury  well  inserted  in  the  steam  pipe.  The  degree  of 
superheating  should  be  taken  as  the  difference  between  the  reading 
of  the  thermometer  for  superheated  steam  and  the  readings  of  the 
same  thermometer  for  saturated  steam  at  the  same  pressure  as 
determined  by  a  special  experiment,  and  not  by  reference  to  steam 
tables. 

For  calculations  relating  to  quality  of  steam  and  corrections  for 
quality  of  steam,  see  Appendices  XVIII  and  XIX. 

XV.  Sampling  the  Coal  and  Determining  its  Moisture.  —  As  each 
barrow  load  or  fresh  portion  of  coal  is  taken  from  the  coal  pile,  a  rep- 
resentative shovelful  is  selected  from  it  and  placed  in  a  barrel  or  box 
in  a  cool  place  and  kept  until  the  end  of  the  trial.  The  samples  are 
then  mixed  and  broken  into  pieces  not  exceeding  one  inch  in  diameter, 
and  reduced  by  the  process  of  repeated  quartering  and  crushing  until 
a  final  sample  weighing  about  five  pounds  is  obtained,  and  the  sizes  of 
the  larger  pieces  are  such  that  they  will  pass  through  a  sieve  with 
J-inch  meshes.  From  this  sample  two  one-quart,  air-tight  glass  pre- 
serving jars,  or  other  air-tight  vessels  which  will  prevent  the  escape  of 
moisture  from  the  sample,  are  to  be  promptly  filled,  and  these  samples 
are  to  be  kept  for  subsequent  determinations  of  moisture  and  of  heat- 
ing value  and  for  chemical  analyses.  During  the  process  of  quar- 
tering, when  the  sample  has  been  reduced  to  about  100  pounds,  a 
quarter  to  a  half  of  it  may  be  taken  for  an  approximate  determination 
of  moisture.  This  may  be  made  by  placing  it  in  a  shallow  iron  pan, 
not  over  three  inches  deep,  carefully  weighing  it,  and  setting  the  pan 
in  the  hottest  place  that  can  be  found  on  the  brickwork  of  the  boiler 


APPENDIX  B  829 

setting  or  flues,  keeping  it  there  for  at  least  12  hours,  and  then 
weighing  it.  The  determination  of  moisture  thus  made  is  believed 
to  be  approximately  accurate  for  anthracite  and  semi-bituminous 
coals,  and  also  for  Pittsburg  or  Youghiogheny  coal;  but  it  cannot  be 
relied  upon  for  coals  mined  west  of  Pittsburg,  or  for  other  coals  con- 
taining inherent  moisture.  For  these  latter  coals  it  is  important  that 
a  more  accurate  method  be  adopted.  The  method  recommended  by 
the  Committee  for  all  accurate  tests,  whatever  the  character  of  the 
coal,  is  described  as  follows: 

Take  one  of  the  samples  contained  in  the  glass  jars  and  subject  it  to 
a  thorough  air-drying  by  spreading  it  in  a  thin  layer  and  exposing  it 
for  several  hours  to  the  atmosphere  of  a  warm  room,  weighing  it 
before  and  after,  thereby  determining  the  quantity  of  surface  moisture 
it  contains.  Then  crush  the  whole  of  it  by  running  it  through  an 
ordinary  coffee  mill  adjusted  so  as  to  produce  somewhat  coarse  grains 
(less  than  -^  inch),  thoroughly  mix  the  crushed  sample,  select  from  it 
a  portion  of  from  10  to  50  grams,  weigh  it  in  a  balance  which  will 
easily  show  a  variation  as  small  as  1  part  in  1000,  and  dry  it  in  an  air 
or  sand  bath  at  a  temperature  between  240  and  280  degrees  F.  for  one 
hour.  Weigh  it  and  record  the  loss,  then  heat  and  weigh  it  again 
repeatedly,  at  intervals  of  an  hour  or  less,  until  the  minimum  weight 
has  been  reached  and  the  weight  begins  to  increase  by  oxidation  of 
a  portion  of  the  coal.  The  difference  between  the  original  and  the 
minimum  weight  is  taken  as  the  moisture  in  the  air-dried  coal.  This 
moisture  test  should  preferably  be  made  on  duplicate  samples,  and  the 
results  should  agree  within  0.3  to  0.4  of  one  per  cent,  the  mean  of 
the  two  determinations  being  taken  as  the  correct  result.  The  sum  of 
the  percentage  of  moisture  thus  found  and  the  percentage  of  surface 
moisture  previously  determined  is  the  total  moisture.  (Appendix  XI.) 

XVI.  Treatment  of  Ashes  and  Refuse.  —  The  ashes  and  refuse  are 
to  be  weighed  in  a  dry  state.     If  it  is  found  desirable  to  show  the 
principal  characteristics  of  the  ash,  a  sample  should  be  subjected  to  a 
proximate  analysis  and  the  actual  amount  of  incombustible  material 
determined.     For  elaborate  trials  a  complete  analysis  of  the  ash  and 
refuse  should  be  made. 

XVII.  Calorific  Tests  and  Analysis  of  Coal.  —  The  quality  of  the 
fuel  should  be  determined  either  by  heat  test  or  by  analysis,  or  by  both. 

The  rational  method  of  determining  the  total  heat  of  combustion  is 
to  burn  the  sample  of  coal  in  an  atmosphere  of  oxygen  gas,  the  coal  to 
be  sampled  as  directed  in  Article  XV  of  this  code.  (See  Appendices 
XIII  and  XIV.) 


830  STEAM  POWER  PLANT  ENGINEERING 

The  chemical  analysis  of  the  coal  should  be  made  only  by  an 
expert  chemist.  The  total  heat  of  combustion  computed  from  the 
results  of  the  ultimate  analysis  may  be  obtained  by  the  use  of  Dulong's 
formula  (with  constants  modified  by  recent  determinations),  viz.: 

14,600  C  +  62,000  f  J3F-—-   +  400°  s>  in  which  C>  H>  °>  and  s  refer 


to  the  proportions  of  carbon,  hydrogen,  oxygen,  and   sulphur  respec- 
tively, as  determined  by  the  ultimate  analysis.* 

It  is  desirable  that  a  proximate  analysis  should  be  made,  thereby 
determining  the  relative  proportions  of  volatile  matter  and  fixed 
carbon.  These  proportions  furnish  an  indication  of  the  leading 
characteristics  of  the  fuel,  and  serve  to  fix  the  class  to  which  it 
belongs.  (Appendix  XII.)  As  an  additional  indication  of  the  char- 
acteristics of  the  fuel,  the  specific  gravity  should  be  determined. 

XVIII.  Analysis  of  Flue  Gases.  —  The  analysis  of  the  flue  gases  is 
an  especially  valuable  method  of  determining  the  relative  value  of 
different    methods   of   firing,    or   of   different    kinds   of   furnaces.     In 
making  these  analyses  great  care  should  be  taken  to  procure  average 
samples,  since  the  composition  is   apt  to  vary  at  different   points   of 
the  flue.     (Appendix  XXXI.)     The  composition  is   also   apt  to  vary 
from   minute   to    minute,  and   for   this   reason   the   drawings  of   gas 
should   last    a    considerable   period    of   time.      Where    complete    de- 
terminations   are    desired,  the  analyses    should    be    intrusted  to  an 
expert  chemist.     For  approximate  determinations  the  Orsatf  or  the 
Hempel  J  apparatus  may  be  used  by  the  engineer.     (See  Appendix 
XXXIII.) 

For  the  continuous  indication  of  the  amount  of  carbonic  acid 
present  in  the  flue  gases,  an  instrument  may  be  employed  which 
shows  the  weight  of  the  sample  of  gas  passing  through  it.  (Appendix 
XXXIX.) 

XIX.  Smoke  Observations.  —  It  is  desirable  to  have  a  uniform  system 
of  determining  and  recording  the  quantity  of  smoke  produced  where 
bituminous  coal  is  used.     The  system  commonly  employed  is  to  express 
the  degree  of  smokiness  by  means  of  percentages  dependent  upon  the 
judgment  of  the  observer.     The  Committee  does  not  place  much  value 

*  Favre  and  Silberman  give  14,544  B.T.U.  per  pound  carbon;  Berthelot  14,647 
B.T.U.  Favre  and  Silberman  give  62,032  B.T.U.  per  pound  hydrogen;  Thomsen 
61,816  B.T.U. 

f  See  R.  S.  Hale's  paper  on  "Flue  Gas  Analysis,"  Transactions,  Vol.  XVIII, 
p.  901. 

J  See  Hempel  on  "Gas  Analysis." 


APPENDIX  B  831 

upon  a  percentage  method,  because  it  depends  so  largely  upon  the  per- 
sonal element,  but  if  this  method  is  used,  it  is  desirable  that,  so  far  as 
possible,  a  definition  be  given  in  explicit  terms  as  to  the  basis  and  method 
employed  in  arriving  at  the  percentage.  The  actual  measurement  of  a 
sample  of  soot  and  smoke  by  some  form  of  meter  is  to  be  preferred. 
(See  Appendices  XXXIV  and  XXXV.) 

XX.  Miscellaneous.  —  In  tests  for  purposes  of  scientific  research,  in 
which  the  determination  of  all  the  variables  entering  into  the  test  is 
desired,   certain  observations  should  be  made  which  are  in  general 
unnecessary  for  ordinary  tests.     These  are  the  measurement  of  the  air 
supply,  the  determination  of  its  contained  moisture,  the  determination 
of  the  amount  of  heat  lost  by  radiation,  of  the  amount  of  infiltration  of 
air  through  the  setting,  and  (by  condensation  of  all  the  steam  made  by 
the  boiler)  of  the  total  heat  imparted  to  the  water. 

As  these  determinations  are  rarely  undertaken,  it  is  not  deemed 
advisable  to  give  directions  for  making  them. 

XXI.  Calculations  of  Efficiency.  —  Two   methods   of   defining  and 
calculating  the  efficiency  of  a  boiler  are  recommended.     They  are: 

1.  Efficiency  of  the  boiler  =  Heat  absorbed  perlb.  combustible  . 

Calorific  value  of  1  Ib.  combustible 

2.  Efficiency  of  the  boiler  and  grate  =  Heat  absorbed  per  Ib.  coal   . 

Calorific  value  of  1  Ib.  coal 

The  first  of  these  is  sometimes  called  the  efficiency  based  on  com- 
bustible, and  the  second  the  efficiency  based  on  coal.  The  first  is 
recommended  as  a  standard  of  comparison  for  all  tests,  and  this  is  the 
one  which  is  understood  to  be  referred  to  when  the  word  "  efficiency  " 
alone  is  used  without  qualification.  The  second,  however,  should  be 
included  in  a  report  of  a  test,  together  with  the  first,  whenever  the 
object  of  the  test  is  to  determine  the  efficiency  of  the  boiler  and  furnace 
together  with  the  grate  (or  mechanical  stoker),  or  to  compare  different 
furnaces,  grates,  fuels,  or  methods  of  firing. 

The  heat  absorbed  per  pound  of  combustible  (or  per  pound  coal)  is 
to  be  calculated  by  multiplying  the  equivalent  evaporation  from  and 
at  212  degrees  per  pound  combustible  (or  coal)  by  965.7  (Appendix 
XX.) 

XXII.  The   Heat  Balance.  —  An  approximate  "  heat  balance,"  or 
statement  of  the  distribution  of  the  heating  value  of  the  coal  among  the 
several  items  of  heat  utilized  and  heat  lost,  may  be  included  in  the 


832 


STEAM  POWER  PLANT  ENGINEERING 


report  of  a  test  when  analyses  of  the  fuel  and  of  the  chimney  gases  have 
been  made.     It  should  be  reported  in  the  following  form: 


HEAT  BALANCE,  OR  DISTRIBUTION  OF  THE  HEATING  VALUE  OF  THE 

COMBUSTIBLE. 

Total  Heat  Value  of  1  pound  of  combustible . .  .  .B.T.U. 


B.T.U. 

Per  Cent. 

1.     Heat  absorbed  by  the  boiler  =  evaporation  from  and  at  212 
degrees  per  pound  of  combustible  X  965.7. 
2.     Loss  due  to  moisture  in  coal  =  per  cent  of  moisture  referred 
to  combustible  -r-  100  X  [(212  -  0  +  966  +  0.48  (T  — 
212)]  (1  =  temperature  of  air  in  the  boiler  room,  T  — 
that  of  the  flue  gases). 
3.     Loss  due  to  moisture  formed  by  the  burning  of  hydrogen 
=  per  cent  of  hydrogen  to  combustible  •*•  100  X  9  X 
[(212  -  0  +  966  +  0.48  (T  -  212)]. 
4.  *  Loss  due  to  heat  carried  away  in  the  dry  chimney  gases  = 
weight  of  gas  per  pound  of  combustible  X  0.24  X 
(T  -  0- 

CO 
5.  f  Loss  due  to  incomplete  combustion  of  carbon  =  —  — 
CU2  +LO 
„  per  cent  C  in  combustible  _ 

100                    •*  HMoO. 

6.     Loss  due  to  unconsumed  hydrogen  and  hydrocarbons,  to 
heating  the  moisture  in  the  air,  to  radiation,  and  un- 
accounted for.     (Some  of  these  losses  may  be  sepa- 
rately itemized  if  data  are  obtained  from  which  they 
may  be  calculated.) 

Totals  

100  00 

*  The  weight  of  gas  per  pound  of  carbon  burned  may  be  calculated  from  the  gas  analyses  as 
follows: 


Dry  gas  per  pound  carbon  = 


11  CO2  +  8  O  +  7  (CO  +  N) 
- 
T  CO) 


in  which  CO2,  CO,  O,  and  N  are 


the  percentages  by  volume  of  the  several  gases.  As  the  sampling  and  analyses  of  the  gases  in  the 
present  state  of  the  art  are  liable  to  considerable  errors,  the  result  of  this  calculation  is  usually 
only  an  approximate  one.  The  heat  balance  itself  is  also  only  approximate  for  this  reason,  as  well 
as  for  the  fact  that  it  is  not  possible  to  determine  accurately  the  percentage  of  unburned  hydrogen 
or  hydrocarbons  in  the  flue  gases.  (See  Appendix  XXXII.) 

The  weight  of  dry  gas  per  pound  of  combustible  is  found  by  multiplying  the  dry  gas  per  pound 
of  carbon  by  the  percentage  of  carbon  in  the  combustible  and  dividing  by  100. 

f  CO2  and  CO  are  respectively  the  percentage  by  volume  of  carbonic  acid  and  carbonic  oxide  in 
the  flue  gases.  The  quantity  10,150  =  No.  heat  units  generated  by  burning  to  carbonic  acid 
one  pound  of  carbon  contained  in  carbonic  oxide. 

XXIII.  Report  of  the  Trial.  —  The  data  and  results  should  be 
reported  in  the  manner  given  in  either  one  of  the  two  following  tables, 
omitting  lines  where  the  tests  have  not  been  made  as  elaborately  as 
provided  for  in  such  tables.  Additional  lines  may  be  added  for  data 
relating  to  the  specific  object  of  the  test.  The  extra  lines  should  be 


APPENDIX  B  833 

classified  under  the  headings  provided  in  the  tables,  and  numbered  as 
per  preceding  line,  with  sub  letters  a,  b,  etc.  The  Short  Form  of  Report, 
Table  No.  2,  is  recommended  for  commercial  tests  and  as  a  convenient 
form  of  abridging  the  longer  form  for  publication  when  saving  of  space 
is  desirable.  For  elaborate  trials,  it  is  recommended  that  the  full  log 
of  the  trial  be  shown  graphically,  by  means  of  a  chart.  (Appendix 
XXXVIII.) 

TABLE  NO.  1. 

DATA  AND  RESULTS  OF  EVAPORATIVE  TEST, 

Arranged  in  accordance  with  the  Complete  Form  advised  by  the  Boiler  Test  Com- 
mittee of  the  American  Society  of  Mechanical  Engineers.     Code  of  1899. 

Made  by of boiler  at to 

determine 

Principal  conditions  governing  the  trial 


Kind  of  fuel  * 

Kind  of  furnace 

State  of  the  weather . 


Method  of  starting  and  stopping  the  test  ("standard"  or  "alternate,"  Art.  X 
and  XI,  Code) 

1.  Date  of  trial 

2.  Duration  of  trial hours. 

Dimensions  and  Proportions. 

A  complete  description  of  the  boiler,  and  drawings  of  the  same  if  of  unusual 
type,  should  be  given  on  an  annexed  sheet.     (See  Appendix  X.) 

3.  Grate  surface width length area square  feet 

4.  Height  of  furnace  . . . inches. 

5.  Approximate  width  of  air  spaces  in  grate inch. 

6.  Proportion  of  air  space  to  whole  grate  surface . .  . per  cent. 

7.  Water-heating  surface square  feet. 

8.  Superheating  surface " 

9.  Ratio  of  water-heating  surface  to  grate  surface —  to  1. 

10.  Ratio  of  minimum  draft  area  to  grate  surface 1  to  — . 

,  •      ^  ^     j  f    fc      r*&r       •   *  .  .  ;- 

Average  Pressures. 

11.  Steam  pressure  by  gauge lb.  per  sq.  in. 

12.  Force  of  draft  between  damper  and  boiler in.  of  water. 

13.  Force  of  draft  in  furnace "  " 

14.  Force  of  draft  or  blast  in  ash  pit "  " 

*  The  items  printed  in  italics  correspond  to  the  items  in  the  "  Short  Form  of 
Code." 


834  STEAM  POWER  PLANT  ENGINEERING 

Average  Temperatures. 

15.  Of  external  air degrees 

16.  Of  fire  room « 

17.  Of  steam « 

18.  Of  feed  water  entering  heater « 

19.  Of  feed  water  entering  economizer « 

20.  Of  feed  water  entering  boiler « 

21.  Of  escaping  gases  from  boiler " 

22.  Of  escaping  gases  from  economizer " 

Fuel. 

23.  Size  and  condition 

24.  Weight  of  wood  used  in  lighting  fire .  . pounds. 

25.  Weight  of  coal  as  fired  * 

26.  Percentage  of  moisture  in  coal  f per  cent. 

27.  Total  weight  of  dry  coal  consumed pounds. 

28.  Total  ash  and  refuse " 

29.  Quality  of  ash  and  refuse 

30.  Total  combustible  consumed .  .  pounds. 

31.  Percentage  of  ash  and  refuse  in  dry  coal per  cent. 

Proximate  Analysis  of  Coal. 

(App.  XII.) 

Of  Coal.  Of  Combustible. 

32.  Fixed  carbon per  cent.  per  cent. 

33    Volatile  matter "  " 

34.  Moisture 

35.  Ash..  "  


100  per  cent.      100  per  cent. 

36.  Sulphur,  separately  determined "  " 

Ultimate  Analysis  of  Dry  Coal. 
(Art.  XVII,  Code.) 

Of  Coal.         Of  Combustible. 

37.  Carbon  (C) per  cent.  per  cent. 

38.  Hydrogen  (H) 

39.  Oxygen  (O) 

40.  Nitrogen  (N) 

41.  Sulphur  (S) 

42.  Ash . .  "  


100  per  cent.      100  per  cent. 
43.  Moisture  in  sample  of  coal  as  received " 

*  Including  equivalent  of  wood  used  in  lighting  the  fire,  not  including  unburned 
coal  withdrawn  from  furnace  at  times  of  cleaning  and  at  end  of  test.  One  pound  of 
wood  is  taken  to  be  equal  to  0.4  pound  of  coal,  or,  in  case  greater  accuracy  is 
desired,  as  having  a  heat  value  equivalent  to  the  evaporation  of  6  pounds  of  water 
from  and  at  212  degrees  per  pound.  (6  X  965.7  =  5794  B.T.U.) 

f  This  is  the  total  moisture  in  the  coal  as  found  by  drying  it  artificially,  as 
described  in  Art.  XV  of  Code. 


APPENDIX  B 


Analysis  of  Ash  and  Refuse. 


835 


44.  Carbon per  cent. 

45.  Earthy  matter 

Fuel  per  Hour. 

46.  Dry  coal  consumed  per  hour pounds. 

47.  Combustible  consumed  per  hour 

48.  Dry  coal  per  square  foot  of  grate  surface  per  hour 

49.  Combustible  per  square  foot  of  water-heating  surface  per  hour.  .  . 

Calorific  Value  of  Fuel. 
(Art.  XVII,  Code.) 

50.  Calorific  value  by  oxygen  calorimeter,  per  pound  of  dry  coal B.T.U. 

51.  Calorific  value  by  oxygen  calorimeter,  per  pound  of  combustible 

52.  Calorific  value  by  analysis,  per  pound  of  dry  coal  * 

53.  Calorific  value  by  analysis,  per  pound  of  combustible 

Quality  of  Steam. 
(App.  XV  to  XIX.) 

54.  Percentage  of  moisture  in  steam per  cent. 

55.  Number  of  degrees  of  superheating degrees. 

56.  Quality  of  steam  (dry  steam  =  unity).     (For  exact  determination 

of  the  factor  of  correction  for  quality  of  steam  see  Appendix 
XVIII) 

Water. 
(App.  I,  IV,  VII,  VIII.) 

57.  Total  weight  of  water  fed  to  boiler  f pounds. 

58.  Equivalent  water  fed  to  boiler  from  and  at  212  degrees 

59.  Water  actually  evaporated,  corrected  for  quality  of  steam 

60.  Factor  of  evaporation  J pounds. 

61.  Equivalent  water  evaporated  into  dry  steam  from  and  at  212 

degrees.     (Item  59  X  Item  60.) " 

Water  per  Hour. 

62.  Water  evaporated  per  hour,  corrected  for  quality  of  steam 

63.  Equivalent  evaporation  per  hour  from  and  at  212  degrees 

64.  Equivalent  evaporation  per  hour  from  and  at  212  degrees  per  square 

foot  of  water-heating  surface 

*  See  formula  for  calorific  value  under  Article  XVII  of  Code. 

f  Corrected  for  inequality  of  water  level  and  of  steam  pressure  at  beginning  and 
end  of  test. 

TT r 

%  Factor  of  evaporation  =  x^-~  in  which  H  and  h  are  respectively  the  total  heat 
9o5.7 

in  steam  of  the  average  observed  pressure,  and  in  water  of  the  average  observed 
temperature  of  the  feed. 


836 


STEAM  POWER  PLANT  ENGINEERING 


Horse  Power. 

65.  Horse  power  developed.     (34£  pounds  of  water  evaporated  per  hour 

into  dry  steam  from  and  at  212  degrees  equals  one  horse  power.)*  . . .  horse 

66.  Builders'  rated  horse  power 


power. 


67.  Percentage  of  builders'  rated  horse  power  developed . 


per  cent. 


Economic  Results. 

68.  Water  apparently  evaporated  per  pound  of  coal  under  actual  condi- 

tions.    (Item  58  -*-  Item  25.) 

69.  Equivalent  evaporation  from  and  at  212  degrees  per  pound  of  wet 

coal.     (Item  61  -f-  Item  25.) 

70.  Equivalent  evaporation  from  and  at  212  degrees  per  pound  of  dry 

coal.     (Item  61  -=-  Item  27.) 

71.  Equivalent  evaporation  from  and  at  212  degrees  per  pound  of  com- 

bustible.    (Item  61  -r-  Item  30.) 

(If  the  equivalent  evaporation,  Items  69,  70,  and  71,  is  not  cor- 
rected for  the  quality  of  steam,  the  fact  should  be  stated.) 


pounds. 


Efficiency. 
(See  Art.  XXI,  Code.) 

72.  Efficiency  of  the  boiler;  heat  absorbed  by  the  boiler  per  pound  of  com- 

bustible divided  by  the  heat  value  of  one  pound  of  combustible  f .  .  .          per  cent. 

73.  Efficiency  of  boiler,  including  the  grate;  heat  absorbed  by  the  boiler, 

per  pound  of  dry  coal  fired,  divided  by  the  heat  value  of  one  pound 
of  dry  coal  J 

Cost  of  Evaporation. 

74.  Cost  of  coal  per  ton  of pounds  delivered  in  boiler  room $ 

75.  Cost  of  fuel  for  evaporating  1000  pounds  of  water  under  observed 

conditions $ 

76.  Cost  of  fuel  used  for  evaporating  1000  pounds  of  water  from  and  at 

212  degrees $ 

Smoke  Observations. 
(App.  XXXIV  and  XXXV.) 

77.  Percentage  of  smoke  as  observed per  cent. 

78.  Weight  of  soot  per  hour  obtained  from  smoke  meter ounces. 

79.  Volume  of  soot  per  hour  obtained  from  smoke  meter cubic  inches. 

*  Held  to  be  the  equivalent  of  30  pounds  of  water  per  hour  evaporated  from  100 
degrees  F.  into  dry  steam  at  70  pounds  gauge  pressure.  (See  Introduction  to  Code.) 

f  In  all  cases  where  the  word  combustible  is  used,  it  means  the  coal  without 
moisture  and  ash,  but  including  all  other  constituents.  It  is  the  same  as  what  is 
called  in  Europe  "coal  dry  and  free  from  ash." 

J  The  heat  value  of  the  coal  is  to  be  determined  either  by  an  oxygen  calorimeter 
or  by  calculation  from  ultimate  analysis. 


APPENDIX  B  837 


Methods  of  Firing. 

80.  Kind  of  firing  (spreading,  alternate,  or  coking) 

81.  Average  thickness  of  fire 

82.  Average  intervals  between  firings  for  each  furnace  during  time 

when  fires  are  in  normal  condition 

83.  Average  interval  between  times  of  levelling  or  breaking  up 


Analyses  of  the  Dry  Gases. 

84.  Carbon  dioxide  (C02) per  cent. 

85.  Oxygen  (O) 

86.  Carbon  monoxide  (CO) 

87.  Hydrogen  and  hydrocarbons 

88.  Nitrogen  (by  difference)  (N) 

100  per  cent. 

TABLE  NO.  2. 
DATA  AND  RESULTS  OP  EVAPORATIVE  TEST, 

Arranged  in  accordance  with  the  Short  Form  advised  by  the  Boiler  Test  Committee 
of  the  American  Society  of  Mechanical  Engineers.     Code  of  1899. 

Made  by on boiler,  at to 

determine 

Kind  of  fuel 

Kind  of  furnace 

Method  of  starting  and  stopping  the  test  (" standard"  or  "alternate,"  Art.    X 

and  XI,  Code) 

Grate  surface square  feet. 

Water-heating  surface " 

Superheating  surface " 

Total  Quantities. 

1.  Date  of  trial 

2.  Duration  of  trial hours. 

3.  Weight  of  coal  as  fired  * pounds. 

4.  Percentage  of  moisture  in  coal  * per  cent. 

5.  Total  weight  of  dry  coal  consumed pounds. 

6.  Total  ash  and  refuse 

7.  Percentage  of  ash  and  refuse  in  dry  coal per  cent. 

8.  Total  weight  of  water  fed  to  the  boiler pounds. 

9.  Water  actually  evaporated,  corrected  for  moisture  or  superheat  in 

steam " 

10.  Equivalent  water  evaporated  into  dry  steam  from  and  at  212 

degrees " 

*  See  foot-notes  of  Complete  Form. 


838 


STEAM   POWER  PLANT  ENGINEERING 


Hourly  Quantities. 

11.  Dry  coal  consumed  per  hour pounds. 

12.  Dry  coal  per  hour  per  square  foot  of  grate  surface " 

13.  Water  fed  per  hour " 

14.  Equivalent  water  evaporated  per  hour  from  and  at  212  degrees 

corrected  for  quality  of  steam " 

15.  Equivalent  water  evaporated  per  hour  per  square  foot  of  water- 

heating  surface " 

•  .        Average  Pressures,  Temperatures,  etc. 

16.  Average  boiler  pressure Ib.  per  sq.  in. 

17.  Average  temperature  of  feed  water degrees. 

18.  Average  temperature  of  escaping  gases " 

19.  Average  force  of  draft  between  damper  and  boiler in.  of  water. 

20.  Percentage  of  moisture  in  steam,  or  number  of  degrees  of  super- 

heating   

Horse  Power. 

21.  Horse  power  developed  (Item  14  -s-  34£) horse  power. 

22.  Builders'  rated  horse  power " 

23.  Percentage  of  builders'  rated  horse  power  developed per  cent. 

Economic  Results. 

24.  Water  apparently  evaporated  per  pound  of  coal  under  actual 

conditions.     (Item  8  -?•  Item  3.) pounds. 

25.  Equivalent  water  actually  evaporated  from  and  at  212  degrees 

per  pound  of  wet  coal.     (Item  9  -f-  Item  3.) " 

26.  Equivalent  evaporation  from  and  at  212  degrees  per  pound  of  dry 

coal.     (Item  9  •*•  Item  5.) 

27.  Equivalent  evaporation  from  and  at  212  degrees  per  pound  of 

combustible.     [Item  9  •*•  (Item  5  -  Item  6).] " 

(If  Items  25,  26,  and  27  are  not  corrected  for  quality  of  steam,  the 
fact  should  be  stated.) 

Efficiency. 

28.  Heating  value  of  the  coal  per  pound B.T.U. 

29.  Heating  value  of  the  combustible  per  pound 

30.  Efficiency  of  boiler  (based  on  combustible) per  cent. 

31.  Efficiency  of  boiler,  including  grate  (based  on  coal) " 

Cost  of  Evaporation. 

32.  Cost  of  coal  per  ton  delivered  in  boiler  room $ 

33.  Cost  of  coal  required  for  evaporation  of   1000  pounds  of  water 

from  and  at  212  degrees $ 


APPENDIX  B  839 

LIST  OF  APPENDICES  TO  CODE  * 

No.  of  Appendix. 

I.  Relative  Weights  of  Water  and  Fuel c.  E.  E. 

II.  Object  of  the  Test.     (I,  1885  Code) j.  c.  H. 

III.  General  Observations.     (II,  1885  Code) c.  T.  P. 

IV.  Precautions  to  be  Observed  in  Making  a  Boiler  Test.     (Ill,   1885 

Code) c.  E.  E. 

V.  Weighing  the  Coal.     (IV,  1885  Code) j.  c.  H. 

VI.  Weighing  the  Coal.      (V,  1885  Code) c.  T.  p. 

VII.  Weighing  the  Water.     (VI,  1885  Code) j.  c.  H. 

VIII.  Measuring  the  Feed  Water.     (VII,  1885  Code) c.  T.  p. 

IX.  Keeping  Time  of  Observations.     (VIII,  1885  Code) j.  c.  H. 

X.  Description  of  Boiler.     (XXIII,  1885  Code) c.  E.  E. 

XI.  Determining  the  Moisture  in  Coal w.  K. 

XII.  Proximate  Analyses  of  Coal w.  K. 

XIII.  Coal  Calorimeter G.  H.  B. 

XIV.  Comparative  Calorimetric  Tests  of  Coals w.  K. 

XV.  Determination  of  the  Moisture  in  the  Steam w.  K. 

XVI.  Correction  for  Radiation  from  Throttling  Calorimeters G.  H.  B. 

XVII.  Combined  Calorimeter  and  Separator G.  H.  B. 

XVIII.  Corrections  for  Quality  of  Steam c.  E.  E. 

XIX.  The  Quality  of  Superheated  Steam G.  H.  B. 

XX.  Efficiency  of  the  Boiler w.  K. 

XXI.  Distribution  of  the  Heating  Value  of  the  Fuel w.  K. 

XXII.  Observation  Blanks.     (Amendment  to  XXIV,  1885  Code) c.  E.  E. 

XXIII.  Horse  Power.     (XXV,  1885  Code) j.  c.  H. 

XXIV.  Steam  Units.     (XXVI,  1885  Code) c.  E.  E. 

XXV.  Discrepancy  between  Commercial  and  Experimental  Results .  .  .  .c.  E.  E. 

XXVI.  Recording  Steam  Gauge.     (IX,  1885  Code) j.  c.  H, 

XXVII.  Pyrometer.     (XIII,  1885  Code) c.  T.  p. 

XXVIII.  Draft  Gauge.     (XIV,  1885  Code) j.  c.  H. 

XXIX.  Draft  Gauge G.  H.  B. 

XXX.  Draft  Gauge w.  K. 

XXXI.  Sampling  Flue  Gases.     (XVI,  1885  Code) j.  c.  H. 

XXXII.  Computation  of  the  Weight  of  Chimney  Gases  from  the  Analysis 

by  Volume  of  Dry  Gas w.  K. 

XXXIII.  The  Orsat  Apparatus  for  Analyzing  Flue  Gases G.  H.  B. 

XXXIV.  Smoke  Measurements G.  H.  B. 

XXXV.  The  Ringelmann  Smoke  Chart w.  K. 

XXXVI.  Starting  and  Stopping  a  Test w.  K. 

XXXVII.  Starting  and  Stopping  a  Test G.  H.  B. 

XXXVIII.  Chart  Showing  Graphically  the  Log  of  a  Trial G.  H.  B. 

XXXIX.  Continuous  Determinations  of  Carbonic  Acid  in  Flue  Gases.  .  .  .G.  H.  B. 

XL.fMeasuring  Radiation  from  Certain  Types  of  Boilers    R.  s.  H. 

XLI.fDetermination  of  the  Moisture  in  Steam  Flowing  through  a 

Horizontal  Pipe D.  s.  j. 

*  Only  a  few  of  the  appendices  are  reprinted. 

t  Contributed  by  members  of  the  Society  and  accepted  by  the  Committee  for 
publication  in  the  Appendix. 


840  STEAM   POWER   PLANT   ENGINEERING 

APPENDIX   XX. 
Efficiency  of  the  Boiler. 

The  efficiency  of  the  boiler,  including  the  grate,  or  the  efficiency  based 
on  coal,  is  the  quotient  arising  from  dividing  the  heat  absorbed  by  the 
boiler  by  the  heating  value  of  the  total  amount  of  coal  supplied  to  the 
boiler,  including  the  coal  which  falls  through  the  grate.  It  may  be 
conveniently  calculated  by  multiplying  the  number  of  pounds  of  water 
evaporated  from  and  at  212  degrees  F.  into  dry  steam  per  pound  of 
dry  coal  by  965.7  and  dividing  the  product  by  the  heating  value  in 
British  thermal  units  of  one  pound  of  dry  coal. 

The  efficiency  of  the  boiler,  not  including  the  grate,  or  the  efficiency 
based  on  combustible,  is  the  quotient  arising  from  dividing  the  heat 
absorbed  by  the  boiler  by  the  heating  value  of  the  combustible  burned. 
It  may  be  calculated  by  multiplying  the  number  of  pounds  of  water 
evaporated  from  and  at  212  degrees  F.  into  dry  steam  per  pound  of 
combustible  by  965.7  and  dividing  the  product  by  the  heating  value  in 
British  thermal  units  of  one  pound  of  combustible;  the  term  "  com- 
bustible "  being  defined  as  coal  dry  and  free  from  ash,  or  the  coal  sup- 
plied to  the  boiler  less  its  moisture  and  the  ash  and  unburned  coal 
which  falls  through  the  grate  or  is  otherwise  withdrawn  from  the  furnace. 

The  efficiency  of  the  boiler,  not  including  the  grate  (or  the  efficiency 
based  upon  combustible),  is  a  more  accurate  measure  of  comparison  of 
different  boilers  than  the  efficiency  including  the  grate  (or  the  efficiency 
based  upon  coal),  for  the  latter  is  subject  to  a  number  of  variable  con- 
ditions, such  as  size  and  character  of  the  coal,  air  spaces  between  the 
grate  bars,  skill  of  the  fireman  in  saving  coal  from  falling  through  the 
grate,  etc.  It  is,  moreover,  subject  to  errors  of  sampling  the  coal  for 
drying  and  for  analysis,  which  affect  the  result  to  a  greater  degree  than 
they  do  the  efficiency  based  upon  combustible,  for  the  reason  that  the 
heating  value  of  one  pound  of  combustible  of  any  sample  selected  from 
a  given  lot,  such  as  a  car  load,  of  coal  is  practically  a  constant  quantity 
and  is  independent  of  the  percentage  of  moisture  and  ash  in  the  sample; 
while  the  sample  itself,  upon  the  heating  value  of  which  the  efficiency 
based  on  coal  is  calculated,  may  differ  in  its  percentage  of  moisture  and 
ash  from  the  average  coal  used  in  the  boiler  test. 

When  the  object  of  a  boiler  test  is  to  determine  its  efficiency  as  an 
absorber  of  heat,  or  to  compare  it  with  other  boilers,  the  efficiency  based 
on  combustible  is  the  one  which  should  be  used;  but  when  the  object  of 
the  test  is  to  determine  the  efficiency  of  the  combination  of  the  boiler, 
the  furnace,  and  the  grate,  the  efficiency  based  on  coal  must  necessarily 
be  used. 


APPENDIX  B  841 

It  has  been  proposed  that  in  reporting  the  efficiency  of  a  boiler  when 
the  fuel  used  contains  hydrogen,  the  efficiency  should  be  considered  to 
be  the  sum  of  the  percentage  of  the  heating  value  of  the  fuel  which  is 
utilized  by  the  boiler  in  making  steam  and  of  the  percentage  of  that 
heating  value  which  is  lost  in  the  shape  of  latent  heat  in  the  moisture  in 
the  chimney  gases,  which  moisture  is  formed  by  the  burning  of  the 
hydrogen.  This  latent  heat  may  amount  to  over  three  per  cent  of  the 
total  heating  value  of  the  fuel.  The  reason  assigned  for  this  proposal 
is  that,  since  it  is  impossible  for  this  heat  to  be  utilized  by  the  boiler 
because  the  gases  are  discharged  at  a  temperature  above  212  degrees  F. 
it  should  not  be  charged  against  the  boiler.  The  writer  does  not  con- 
sider it  advisable  that  this  method  of  reporting  the  efficiency  should 
be  adopted  (1)  because  it  is  opposed  to  the  generally  accepted  definition 
of  efficiency,  which  is  the  useful  work  received  from  an  apparatus 
divided  by  the  work  (or  heat  value  of  the  fuel)  put  into  it;  (2)  because 
in  order  to  calculate  it  it  is  necessary  to  know  both  the  percentage  of 
hydrogen  in  the  coal  and  whether  or  not  all  of  this  hydrogen  has  been 
burned  to  H20,  the  first  requiring  an  analysis  of  the  coal,  which  is  not 
always  obtainable,  and  the  second  an  analysis  of  the  gases  for  hydrogen, 
which  cannot  be  obtained  with  any  approximation  to  accuracy  with  our 
present  methods  of  sampling  and  analyzing  gases;  and  (3)  because  it 
is  opposed  to  the  almost  universal  custom  in  reporting  boiler  tests.  It 
is  true  that  the  latent  heat  of  the  H2O  in  the  chimney  gases  cannot  be 
utilized  (unless  an  economizer  which  discharges  its  gases  below  212 
degrees  is  used),  and  it  is  not  the  fault  of  the  boiler  that  it  cannot  be 
utilized.  It  may  be  considered  the  misfortune  of  the  boiler,  when  tested 
with  hydrogenous  coal,  similar  to  the  misfortune  under  which  an  engine 
labors  when  it  is  tested  while  supplied  with  a  condenser  which  gives  a 
vacuum  of  less  than  30  inches  of  mercury.  The  engine  might  give  a 
higher  efficiency  with  a  vacuum  of  30  inches  than  it  would  with  one  of 
27  or  28  inches;  but  it  is  not  customary  to  credit  the  engine  with  the 
efficiency  which  it  loses  on  account  of  the  imperfect  vacuum. 

Since  it  is  well  understood  that  a  boiler  cannot  show  quite  as  high  an 
efficiency  (as  commonly  defined)  when  using  bituminous  coal  high  in 
hydrogen  as  when  using  anthracite  nearly  free  from  hydrogen,  no  harm 
is  done,  and  much  confusion  is  avoided,  by  reporting  the  efficiency  as 
the  percentage  of  the  heating  value  of  the  coal  which  is  actually  utilized 
in  making  steam.  The  fact  that  bituminous  coal  is  used  is  always 
stated  in  the  report  of  a  test  made  with  that  coal.  If  desired  a  state- 
ment may  also  be  made  in  the  "  heat  balance"  of  the  approximate  or  esti- 
mated percentage  of  heat  which  is  lost  in  the  latent  heat  of  the  moisture 
in  the  chimney  gases,  together  with  the  loss  due  to  moisture  in  the  coal. 


842  STEAM  POWER  PLANT  ENGINEERING 

APPENDIX   XXV. 
Discrepancy  between  Commercial  and  Experimental  Results. 

The  final  result  sought  by  manufacturers,  in  initiating  tests  of  steam 
or  other  machinery  in  actual  use,  is  the  value  of  the  work  done  measured 
in  dollars  and  cents.  In  some  cases  the  broad  question  is  raised  as  to 
the  saving  that  may  be  accomplished  by  installing  improved  boilers, 
engines,  or  other  machinery;  but  more  generally  it  is  desired  to  ascer- 
tain what  can  be  done  to  produce  saving  with  the  apparatus  already  in 
place  under  the  actual  conditions  that  prevail  at  the  particular  location. 
In  both  these  cases  it  is  necessary  to  ascertain  the  average  cost  of  the 
work  done  commercially  previous  to  the  test.  Frequently,  in  fact 
generally,  this  important  fact  will  not  be  'ascertained  by  an  elaborate 
trial,  for  the  reason  that  everything  will  be  put  in  order  for  the  test, 
and  all  details  of  the  trial  be  conducted  so  carefully  that  the  losses  due 
to  average  carelessness  or  want  of  skill  in  the  past  will  be  eliminated, 
the  engineer  making  the  test  will  not  receive  proper  credit,  and  the 
owners  on  seeing  the  report  may  conclude  that  they  are  already  doing 
very  well,  and  perhaps  continue  old  methods  with  fancied  security.  If 
the  cost  of  the  output  of  the  factory  for  a  given  time  were  ascertained 
in  terms  of  the  coal  burned  during  the  same  time,  and  compared  with 
the  corresponding  cost  for  the  time  of  the  trial,  the  latter  would  fre- 
quently be  found  to  be  one-eighth  to  one-third  less  than  the  former,  and 
it  might  not  be  possible  to  tell  what  had  caused  the  difference;  for 
instance,  whether  it  was  due  to  putting  in  order  the  machinery  prior  to 
the  tests,  to  greater  care  exercised  by  the  fireman  under  the  spur  of 
careful  watching,  or  whether,  as  is  usually  claimed,  the  coal  was  different, 
etc.,  etc.  The  losses  are  generally  due  in  the  main  to  the  carelessness  of 
the  firemen.  It  follows,  therefore,  that  the  cost  of  the  power  under 
average  conditions  must  be  obtained  in  some  quiet  way  preliminarily. 
Frequently  the  comparison  of  the  output  of  the  factory  with  the  coal 
burned  will  not  be  sufficiently  accurate,  and  it  will  be  necessary  to 
devise  some  corresponding  check  which  will  not  interfere  with  the 
regular  routine  of  the  establishment.  The  work  of  the  boilers  may  be 
checked  by  arranging  a  meter  so  as  to  continuously  measure  the  feed 
water;  and  its  record,  compared  with  the  total  weight  of  coal  purchased, 
will  frequently  give  the  check  desired.  Such  a  check  becomes  more 
difficult  when  it  is  desirable  to  ascertain  the  performances  of  particular 
boilers,  and  the  coal  supply  is  common  to  all  boilers;  but  by  assigning 
particular  weighed  car  loads  of  coal  to  the  particular  boilers,  without 
any  intimation  to  the  firemen  that  they  are  being  watched,  it  may  be 


APPENDIX  B  843 

possible  to  ascertain  the  average  performance  of  the  boilers  used  for 
the  particular  purpose.  Preliminary  experiments  of  this  kind  conducted 
without  notice  to  employees,  and  continued  through  a  long  period,  will 
furnish  a  basis  for  comparison  with  elaborate  tests,  and  it  will  then  be 
possible  to  point  out  clearly  where  the  several  losses  have  taken  place, 
and  the  testing  engineer  will  get  the  credit  for  the  saving  shown. 

C.  E.  E. 

APPENDIX   XXXV. 
The  Ringelmann  Smoke  Chart. 

Professor  Ringelmann,  of  Paris,  has  invented  a  system  of  determining 
the  relative  density  or  blackness  of  smoke,  which  has  been  communi- 
cated to  the  writer  by  Mr.  Bryan  Donkin,  of  London,  and  published  in 
Engineering  News  of  November  11,  1897.  In  making  observations  of 
the  smoke  proceeding  from  a  chimney  four  cards  ruled  like  those  in  the 
cut,  together  with  a  card  printed  in  solid  black  and  another  left  entirely 
white,  are  placed  in  a  horizontal  row  and  hung  at  a  point  about  50  feet 
from  the  observer  and  as  nearly  as  convenient  in  line  with  the  chimney. 
At  this  distance  the  lines  become  invisible,  and  the  cards  appear  to  be 
of  different  shades  of  gray,  ranging  from  very  light  gray  to  almost  black. 
The  observer  glances  from  the  smoke  coming  from  the  chimney  to  the 
cards,  which  are  numbered  from  0  to  5,  determines  which  card  most 
nearly  corresponds  with  the  color  of  the  smoke  and  makes  a  record 
accordingly,  noting  the  time.  Observations  should  be  made  continu- 
ously during  say.  one  minute,  and  the  estimated  average  density  during 
that  minute  recorded,  and  so  on,  records  being  made  once  every  minute. 
The  average  of  all  the  records  made  during  a  boiler  test  is  taken  as  the 
average  figure  for  the  smoke  density  during  the  test,  and  the  whole  of 
the  record  is  plotted  on  cross-section  paper  in  order  to  show  how  the 
smoke  varied  in  density  from  time  to  time.  A  rule  by  which  the  cards 
may  be  reproduced  is  given  by  Professor  Ringelmann  as  follows : 

Card  0  —  All  white. 

Card  1  —  Black  lines  1  mm.  thick,  10  mm.  apart,  leaving  spaces  9  mm. 
square. 

Card  2  —  Lines  2.3  mm.  thick,  spaces  7.7  mm.  square. 
Card  3  —  Lines  3.7  mm.  thick,  spaces  6.3  mm.  square. 
Card  4  —  Lines  5.5  mm.  thick,  spaces  4.5  mm.  square. 
Card  5  —  All  black. 

The  cards  as  printed  in  Fig.  452  are  much  smaller  than  those  used  by 
Professor  Ringelmann,  but  the  thickness  and  the  spacing  of  the  lines 
.are  the  same.  w.  *K 


844  STEAM  POWER  PLANT  ENGINEERING 


No.l 


No.2 


No.3  No.4 

Flo.  452.    Ringelman  Smoke  Chart. 


APPENDIX  B 


845 


APPENDIX   XXXVIII. 
Chart  Showing  Graphically  the  Log  of  a  Trial. 

The  well-known  method  of  plotting  observations  and  data  on  cross- 
section  paper  and  making  a  chart  applying  to  the  test  is  a  useful 
means  of  representing  the  exact  uniformity  of  conditions  existing 

Lb-  CHART  SHOWING  LOG'OF  BOILER  TEST. 


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during  a  trial.  Such  a  chart  is  illustrated  in  Fig.  453,  in  which  the 
abscissae  represent  times  and  the  ordinates,  on  appropriate  scales, 
the  various  observations  and  data.  G.  H.  B. 


APPENDIX  C. 

RULES  FOR  CONDUCTING  STEAM-ENGINE  TESTS  * 
Code  of  1902. 

I.  Object  of  Test.  —  Ascertain  at  the  outset  the  specific  object  of 
the  test,  whether  it  be  to  determine  the  fulfillment  of  a  contract  guar- 
antee, to  ascertain  the  highest  economy  obtainable,  to  find  the  working 
economy  and  defects  under  conditions  as  they  exist,  to  ascertain  the 
performance    under    special    conditions,    to    determine    the    effect    of 
changes  in  the  conditions,  or  to  find  the  performance  of  the  entire 
boiler  and  engine  plant,  and  prepare  for  the  test  accordingly. 

II.  General  Condition  of  the  Plant.  —  Examine  the  engine  and  the 
entire  plant  concerned  in  the  test;  note  its  general  condition  and  any 
points  of  design,  construction,  or  operation  which  bear  on  the  objects 
in  view.     Make  a  special  examination  of  the  valves  and  pistons  for 
leakage  by  applying  the  working  pressures  with  the  engine  at  rest, 
and  observe  the  quantity  of  steam,  if  any,  blowing  through  per  hour. 

If  the  trial  has  for  an  object  the  determination  of  the  highest 
efficiency  obtainable,  the  valves  and  pistons  must  first  be  made  tight, 
and  all  parts  of  the  engine  and  its  auxiliaries,  and  all  other  parts  of  the 
plant  concerned,  should  be  put  in  the  best  possible  working  condition. 

III.  Dimensions,  etc.  —  Measure   or   check  the  dimensions   of   the 
cylinders  in  any  case,  this  being  done  when  they  are  hot.     If  they  are 
much  worn,   the  average  diameter  should  be  determined.     Measure 
also  the  clearance,  which  should  be  done,  if  possible,  by  filling  the 
spaces  with  water  previously  measured,  the  piston  being  placed  at  the 
end  of  the  stroke.     If  the  clearance  cannot  be  measured  directly,  it 
can  be  determined  approximately  from  the  working  drawings  of  the 
cylinder. 

Measure  also  the  dimensions  of  auxiliaries  and  accessories,  also  those 
of  the  boilers  so  far  as  concerned  in  attaining  the  objects.  It  is  well 
to  supplement  these  determinations  with  a  sketch  or  sketches  showing 
the  general  features  and  arrangement  of  the  different  parts  of  the  plant. 

*  From  the  final  report  of  the  committee  appointed  to  standardize  a  system  of 
testing  steam  engines.  Trans.  A.S.M.E.,  Vol.  XXIV.  Greatly  abridged. 

846 


APPENDIX  C  847 

IV.  Coal  —  When  the  trial  involves  the  complete  plant,  embracing 
boilers  as  well  as  engine,  determine  the  character  of  coal  to  be  used. 
The  class,  name  of  the  mine,  size,  moisture,  and  quality  of  the  coal 
should  be  stated  in  the  report.     It  is  desirable,  for  purposes  of  com- 
parison, that  the  coal  should  be  of  some  recognized  standard  quality 
for  the  locality  where  the  plant  is  situated. 

V.  Calibration    of  Instruments.  —  All    instruments    and    apparatus 
should  be   calibrated   and  their  reliability   and  accuracy   verified  by 
comparison  with  recognized  standards.     Such  apparatus  as  is  liable  to 
change  or  become  broken  during  a  test,  as  gauges,  indicator  springs, 
and  thermometers,   should   be   calibrated  before   and   after  the  test. 
The  accuracy  of  scales  should  be  verified  by  standard  weights.     When 
a  water  meter  is  used,  special  attention  should  be  given  to  its  cali- 
bration, verifying  it  both  before  and  after  the  trial,  and,  if  possible, 
during  its  progress,  the  conditions  in  regard  to  water  pressure  and 
rate  of  flow  being  made  the  same  in  the  calibrations  as  exist  through- 
out the  trial. 

VI.  Leakages  of  Steam,  Water,  etc.  —  In  all  tests  except  those  of  a 
complete  plant  made  under  conditions  as  they  exist,  the  boiler  and  its 
connections,  both  steam  and  feed,  as  also  the  steam  piping  leading  to 
the  engine  and  its  connections,  should,  so  far  as  possible,  be  made 
tight.     If  absolute  tightness  cannot  be  obtained  (in  point  of  fact  it 
rarely  can  be),  proper  allowance  should  be  made  for  such  leakage  in 
determining  the  steam  actually  consumed  by  the  engine.     This,  how- 
ever, is  not  required  where  a  surface  condenser  is  used  and  the  water 
consumption  is   determined  by  measuring  the   discharge  of   the    air 
pump.     In  such  cases  it  is  necessary  to  make  sure  that  the  condenser 
is  tight,  both  before  and  after  the  test,  against  the  entrance  of  circu- 
lating water,  or  if  such  occurs  to  make  proper  correction  for  it,  deter- 
mining it   under  the  working  difference  of    pressure.      Should  there 
be  excessive  leakage  of    the  condenser  it  should  be  remedied  before 
the    test    is    made.       When   the    steam    consumption  is    determined 
by    measuring    the  discharge    of  the    air  pump,    any  leakage   about 
the  valve  or  piston  rods  of  the  engine  should  be  carefully  guarded 
against. 

Make  sure  that  there  is  no  leakage  at  any  of  the  connections  with 
the  apparatus  provided  for  measuring  and  supplying  the  feed  water 
which  could  affect  the  results.  All  connections  should,  so  far  as  pos- 
sible, be  visible  and  be  blanked  off,  and  where  this  cannot  be  done, 
satisfactory  assurance  should  be  obtained  that  there  is  no  leakage 
either  in  or  out. 


848  STEAM   POWER  PLANT  ENGINEERING 

VII.  Duration  of  Test.  —  The   duration  of   a  test  should   depend 
largely  upon  its  character  and  the  objects  in  view.     The  standard  heat 
test  of  an  engine,  and,  likewise,  a  test  for  the  simple  determination  of 
the  feed-water  consumption,   should  be   continued  for  at  least   five 
hours,  unless  the  class  of  service  precludes  a  continuous  run  of  so  long 
duration.     It  is  desirable  to  prolong  the  test  the  number  of  hours 
stated  to  obtain  a  number  of  consecutive  hourly  records  as  a  guide  in 
analyzing  the  reliability  of  the  whole. 

Where  the  water  discharged  from  the  surface  condenser  is  measured 
for  successive  short  intervals  of  time,  and  the  rate  is  found  to  be  uni- 
form, the  test  may  be  of  a  much  shorter  duration  than  where  the  feed 
water  is  measured  to  the  boiler.  The  longer  the  test  with  a  given  set 
of  conditions  the  more  accurate  the  work,  and  no  test  should  be  so 
short  that  it  cannot  be  divided  into  several  intervals  which  will  give 
results  agreeing  substantially  with  each  other. 

The  commercial  test  of  a  complete  plant,  embracing  boilers  as  well 
as  engine,  should  continue  at  least  one  full  day  of  twenty-four  hours, 
whether  the  engine  is  in  motion  during  the  entire  time  or  not.  A 
continuous  coal  test  of  a  boiler  and  engine  should  be  of  at  least  ten 
hours'  duration,  or  the  nearest  multiple  of  the  interval  between  times 
of  cleaning  fires. 

VIII.  Starting  and  Stopping  a  Test.  —  (a)  Standard  Heat  Test  and 
Feed-Water  Test  of  Engine  :  The  engine  having  been  brought  to  the 
normal  condition  of  running,  and  operated  a  sufficient  length  of  time 
to  be  thoroughly  heated  in  all  its  parts,  and  the  measuring  apparatus 
having  been  adjusted  and  set  to  work,  the  height  of  water  in  the 
gauge  glasses  of  the  boilers  is  observed,  the  depth  of  water  in  the 
reservoir  from  which  the  feed  water  is  supplied  is  noted,  the  exact 
time  of  day  is  observed,  and  the  test  held  to  commence.     Thereafter 
the  measurements  determined  upon  for  the  test  are  begun  and  carried 
forward  until  its  close.     If  practicable,  the  test  may  be  commenced  at 
some  even  hour  or  minute,  but  it  is  of  the  first  importance  to  begin  at 
such  time  as  reliable  observations  of  the  water  heights  are  obtained, 
whatever  the  exact  time  happens  to  be  when  these  are  satisfactorily 
determined.     When  the  time  for  the  close  of  the  test   arrives,   the 
water  should,  if  possible,  be  brought  to  the  same  height  in  the  glasses 
and  to  the  same  depth  in  the  feed-water  reservoir  as  at  the  beginning, 
delaying  the  conclusion  of  the  test  if  necessary  to  bring  about  this 
similarity  of  conditions.     If  differences  occur,  the  proper  corrections 
must  be  made. 

(fe)  Complete  Engine  and  Boiler  Test :  For  a  continuous  running 
test  of  combined  engine  or  engines,  and  boiler  or  boilers,  the  same 


APPENDIX  C  849 

directions  apply  for  beginning  and  ending  the  feed-water  measurements 
as  those  just  referred  to  under  Section  (a).  The  time  of  beginning  and 
ending  such  a  test  should  be  the  regular  time  of  cleaning  the  fires,  and 
the  exact  time  of  beginning  and  ending  should  be  the  time  when  the 
fires  are  fully  cleaned,  just  preparatory  to  putting  on  fresh  coal.  In  cases 
where  there  are  a  number  of  boilers,  and  it  is  inconvenient  or  undesirable 
to  clean  all  fires  at  once,  the  time  of  beginning  the  test  should  be  deferred 
until  they  are  all  cleaned  and  in  a  satisfactory  state,  all  the  fires  being 
then  burned  down  to  a  uniformly  thin  condition,  the  thickness  and 
condition  being  estimated  and  the  test  begun  just  before  firing  the  new 
coal  previously  weighed.  The  ending  of  the  test  is  likewise  deferred 
until  the  fires  are  all  satisfactorily  cleaned,  being  again  burned  down  to 
the  same  uniformly  thin  condition  as  before,  and  the  time  of  closing 
being  taken  just  before  replenishing  the  fires  with  new  coal. 

For  a  commercial  test  of  a  combined  engine  and  boiler,  whether  the 
engine  runs  continuously  for  the  full  twenty-four  hours  of  the  day  or 
only  a  portion  of  the  time,  the  fires  in  the  boilers  being  banked  during 
the  time  when  the  engine  is  not  in  motion,  the  beginning  and  ending  of 
the  test  should  occur  at  the  regular  time  of  cleaning  the  fires,  the  method 
followed  being  that  already  given.  In  cases  where  the  engine  is  not  in 
continuous  motion,  as,  for  example,  in  textile  mills,  where  the  working 
time  is  ten  or  eleven  hours  out  of  the  twenty-four,  and  the  fires  are 
cleaned  and  banked  at  the  close  of  the  day's  work,  the  best  time  for 
starting  and  stopping  a  test  is  the  time  just  before  banking,  when  the 
fires  are  well  burned  down  and  the  thickness  and  condition  can  be  most 
satisfactorily  judged.  In  these,  as  in  all  other  cases  noted,  the  test 
should  be  begun  by  observing  the  exact  time,  the  thickness  and  condi- 
tion of  the  fires  on  the  grates,  the  height  of  water  in  the  gauge  glasses  of 
the  boilers,  the  depth  of  the  water  in  the  reservoir  from  which  the  feed 
water  is  supplied,  and  other  conditions  relating  to  the  trial,  the  same 
observations  being  again  taken  at  the  end  of  the  test,  and  the  conditions 
in  all  respects  being  made  as  nearly  as  possible  the  same  as  at  the 
beginning. 

IX.  Measurement  of  Heat  Units  consumed  by  the  Engine.  —  The 
measurement  of  the  heat  consumption  requires  the  measurement  of 
each  supply  of  feed  water  to  the  boiler  —  that  is,  the  water  supplied  by 
the  main  feed  pump,  that  supplied  by  auxiliary  pumps,  such  as  jacket 
water,  water  from  separators,  drips,  etc.,  and  water  supplied  by  gravity 
or  other  means;  also  the  determination  of  the  temperature  of  the  water 
supplied  from  each  source,  together  with  the  pressure  and  quality  of 
the  steam. 

The  temperatures  at  the  various  points  should  be  those  applying  to 


850  STEAM   POWER  PLANT  ENGINEERING 

the  working  conditions.  The  temperature  of  the  feed  water  should  be 
taken  near  the  boiler.  This  causes  the  engine  to  suffer  a  disadvantage 
from  the  heat  lost  by  radiation  from  the  pipes  which  carry  the  water 
to  the  boiler,  but  it  is,  nevertheless,  advisable  on  the  score  of  simplicity. 
Such  pipes  would,  therefore,  be  considered  a  portion  of  the  engine  plant. 
This  conforms  with  the  rule  already  recommended  for  the  tests  of 
pumping  engines  where  the  duty  per  million  heat  units  is  computed 
from  the  temperature  of  the  feed  water  taken  near  the  boiler.  It 
frequently  happens  that  the  measurement  of  the  water  requires  a  change 
in  the  usual  temperature  of  supply.  For  example,  where  the  main 
supply  is  ordinarily  drawn  from  a  hot  well  in  which  the  temperature  is 
say  100  degrees  F.,  it  may  be  necessary,  owing  to  the  low  level  of 
the  well,  to  take  the  supply  from  some  source  under  a  pressure  or  head 
sufficient  to  fill  the  weighing  tanks  used,  and  this  supply  may  have  a 
temperature  much  below  that  of  the  hot  well;  possibly  as  low  as  40 
degrees  F.  The  temperature  to  be  used  is  not  the  temperature  of 
the  water  as  weighed  in  this  case,  but  that  of  the  working  temperature 
of  the  hot  well.  The  working  temperature  in  cases  like  this  must  be 
determined  by  a  special  test,  and  included  in  the  log  sheets. 

The  heat  to  be  determined  is  that  used  by  the  entire  engine  equip- 
ment, embracing  the  main  cylinders  and  all  auxiliary  cylinders  and 
mechanism  concerned  in  the  operation  of  the  engine,  including  the  air 
pump,  circulating  pump,  and  feed  pumps,  also  the  jacket  and  reheater 
when  these  are  used.  No  deduction  is  to  be  made  for  steam  used  by 
auxiliaries  unless  these  are  shown  by  test  to  be  unduly  wasteful.  In 
this  matter  an  exception  should  be  made  in  cases  of  guarantee  tests 
where  the  engine  contractor  furnishes  all  the  auxiliaries  referred  to. 
He  should,  in  that  case,  be  responsible  for  the  whole,  and  no  allowance 
should  be  made  for  inferior  economy,  if  such  exists.  Should  a  deduction 
be  made  on  account  of  the  auxiliaries  being  unduly  wasteful,  the  method 
of  waste  and  its  extent,  as  compared  with  the  wastes  of  the  main  engine 
or  other  standard  of  known  value,  shall  be  reported  definitely. 

The  steam  pressure  and  the  quality  of  the  steam  are  to  be  taken  at 
some  point  conveniently  near  the  throttle  valve.  The  quantity  of 
steam  used  by  the  calorimeter  must  be  determined  and  properly  allowed 
for.  (See  Article  XVI,  on  "  Quality  of  Steam.") 

X.  Measurement  of  Feed  Water  or  Steam  Consumption  of  Engine, 
etc.  —  The  method  of  determining  the  steam  consumption  applicable  to 
all  plants  is  to  measure  all  the  feed  water  supplied  to  the  boilers,  and 
deduct  therefrom  the  water  discharged  by  separators  and  drips,  as  also 
the  water  and  steam  which  escape  on  account  of  leakage  of  the  boiler 
and  its  pipe  connections  and  leakage  of  the  steam  main  and  branches 


APPENDIX  C  851 

connecting  the  boiler  and  the  engine.  In  plants  where  the  engine 
exhausts  into  a  surface  condenser  the  steam  consumption  can  be 
measured  by  determining  the  quantity  of  water  discharged  by  the  air 
pump,  corrected  for  any  leakage  of  the  condenser,  and  adding  thereto 
the  steam  used  by  jackets,  reheaters,  and  auxiliaries  as  determined 
independently.  If  the  leakage  of  the  condenser  is  too  large  to  satis- 
factorily allow  for  it,  the  condenser  should,  of  course,  be  repaired  and 
the  leakage  again  determined  before  making  the  test. 

In  measuring  the  water  it  is  best  to  carry  it  through  a  tank  or  tanks 
resting  on  platform  weighing  scales  suitably  arranged  for  the  purpose, 
the  water  being  afterwards  emptied  into  a  reservoir  beneath,  from  which 
the  pump  is  supplied. 

Where  extremely  large  quantities  of  water  must  be  measured,  or  in 
some  places  relatively  small  quantities,  the  orifice  method  of  measuring 
is  one  that  can  be  applied  with  satisfactory  results.  In  this  case  the 
average  head  of  water  on  the  orifice  must  be  determined,  and,  further- 
more, it  is  important  that  means  should  be  at  hand  for  calibrating  the 
discharge  of  the  orifice  under  the  conditions  of  use. 

The  corrections  or  deductions  to  be  made  for  leakage  above  referred 
to  should  be  applied  only  to  the  standard  heat-unit  test  and  tests  for 
determining  simply  the  steam  or  feed-water  consumption,  and  not  to 
coal  tests  of  combined  engine  and  boiler  equipment.  In  the  latter,  no 
correction  should  be  made  except  for  leakage  of  valves  connecting  to 
other  engines  and  boilers,  or  for  steam  used  for  purposes  other  than  the 
operation  of  the  plant  under  test.  Losses  of  heat  due  to  imperfections 
of  the  plant  should  be  charged  to  the  plant,  and  only  such  losses  as  are 
concerned  in  the  working  of  the  engine  alone  should  be  charged  to  the 
engine. 

In  measuring  jacket  water  or  any  supply  under  pressure  which  has 
a  temperature  exceeding  212  degrees  F.,  the  water  should  first  be 
cooled,  as  may  be  done  by  discharging  it  into  a  tank  of  cold  water 
previously  weighed,  or  by  passing  it  through  a  coil  of  pipe  submerged 
in  running  and  colder  water,  preventing  thereby  the  loss  of  evapo- 
ration which  occurs  when  such  hot  water  is  discharged  into  the 
open  air. 

XI.  Measurement  of  Steam  used  by  Auxiliaries.  —  Although  the 
steam  used  by  the  auxiliaries  —  embracing  the  air  pump,  circulating 
pump,  feed  pump,  and  any  other  apparatus  of  this  nature,  supposing 
them  to  be  steam-driven,  also  the  steam  jackets,  reheaters,  etc.,  which 
consume  steam  required  for  the  operation  of  the  engine  —  is  all 
included  in  the  measurement  of  the  steam  consumption,  as  pointed 
out  in  Article  X,  yet  it  is  highly  desirable  that  the  quantity  of  steam 


852  STEAM   POWER  PLANT  ENGINEERING 

used  by  the  auxiliaries,  and  in  many  cases  that  used  by  each  auxiliary, 
should  be  determined  exactly,  so  that  the  net  consumption  of  the 
main  engine  cylinders  may  be  ascertained  and  a  complete  analysis 
made  of  the  entire  work  of  the  engine  plant.  Where  the  auxiliary 
cylinders  are  non-condensing,  the  steam  consumption  can  often  be 
measured  by  carrying  the  exhaust  for  the  purpose  into  a  tank  of  cold 
water  resting  on  scales  or  through  a  coil  of  pipe  surrounded  by  cold 
running  water.  Another  method  is  to  run  the  auxiliaries  as  a  whole, 
or  one  by  one,  from  a  spare  boiler  (preferably  a  small  vertical  one), 
and  measure  the  feed  water  supplied  to  this  boiler.  The  steam  used 
by  the  air  and  circulating  pumps  may  be  measured  by  running  them 
under,  as  near  as  possible,  the  working  conditions  and  speed,  the  main 
engine  and  other  auxiliaries  being  stopped,  and  testing  the  con- 
sumption by  the  measuring  apparatus  used  on  the  main  trial.  For  a 
short  trial,  to  obtain  approximate  results,  measurement  can  be  made 
by  the  water-gauge  glass  method,  the  feed  supply  being  shut  off. 
When  the  engine  has  a  surface  condenser,  the  quantity  of  steam  used 
by  the  auxiliaries  may  be  ascertained  by  allowing  the  engine  alone  to 
exhaust  into  the  condenser,  measuring  the  feed  water  supplied  to  the 
boiler  and  the  water  discharged  by  the  air  pump,  and  subtracting  one 
from  the  other,  after  allowing  for  losses  by  leakage. 

XII.  Coal  Measurement.  —  (a)  Commercial  Tests :  In  commercial 
tests  of  the  combined  engine  and  boiler  equipment,  or  those  made 
under  ordinary  conditions  of  commercial  service,  the  test  should,  as 
pointed  out  in  Article  VII,  extend  over  the  entire  period  of  the  day; 
that  is,  twenty-four  hours,  or  a  number  of  days  of  that  duration. 
Consequently,  the  coal  consumption  should  be  determined  for  the 
entire  time.  If  the  engine  runs  but  a  part  of  the  time,  and  during  the 
remaining  portion  the  fires  are  banked,  the  measurement  of  coal  should 
include  that  used  for  banking.  It  is  well,  however,  in  such  cases,  to 
determine  separately  the  amount  consumed  during  the  time  the 
engine  is  in  operation  and  that  consumed  during  the  period  while  the 
fires  are  banked,  so  as  to  have  complete  data  for  purposes  of  analysis 
and  comparison,  using  suitable  precautions  to  obtain  reliable  measure- 
ments. The  measurement  of  coal  begins  with  the  first  firing,  after 
cleaning  the  furnaces  and  burning  down  at  the  beginning  of  the  test, 
as  pointed  out  in  Article  VIII,  and  ends  with  the  last  firing,  at  the 
expiration  of  the  allotted  time. 

(6)  Continuous  Running  Tests :  In  continuous  running  tests  which, 
as  pointed  out  in  Article  VII,  cover  one  or  more  periods  which  elapse 
between  the  cleaning  of  the  fires,  the  same  principle  applies  as  that 
mentioned  under  the  above  heading  (a);  viz.,  the  coal  measurement 


APPENDIX  C  853 

begins  with  the  first  firing,  after  cleaning  and  burning  down,  and  the 
measurement  ends  with  the  last  firing,  before  cleaning  and  burning 
down  at  the  close  of  the  trial. 

(c)  Coal  Tests  in  General :    When  not    otherwise   specially  under- 
stood, a  coal  test  of  a  combined  engine  and  boiler  plant  is  held  to  refer 
to  the  commercial  test  above  noted,   and  the  measurement  of  coal 
should  conform  thereto. 

In  connection  with  coal  measurements,  whatever  the  class  of  tests, 
it  is  important  to  ascertain  the  percentage  of  moisture  in  the  coal, 
the  weight  of  ashes  and  refuse,  and,  where  possible,  the  approximate 
and  ultimate  analysis  of  the  coal,  following  all  the  methods  and  details 
advocated  in  the  latest  report  of  the  Boiler  Test  Committee  of  the 
Society.  (See  Vol.  XXI,  p.  34.)  • 

(d)  Other  Fuels  than  Coal :   For  all  other  solid  fuels  than  coal  the 
same  directions  in  regard  to  measurement  should  be  followed  as  those 
given  for  coal.     If  the  boilers  are  run  with  oil  or  gas,  the  measure- 
ments relating  to  stopping  and  starting  are  much  simplified,  because 
the  fuel  is  burned  as  fast  as  supplied,  and  there  is  no  body  of  fuel  con- 
stantly in  the  furnace,  as  in  the  case  of  using  solid  fuel.     When  oil 
is  used,  it  should  be  weighed,  and  when  gas  is  used,  it  should  be 
measured  in  a  calibrated  gas  meter  or  a  gasometer. 

XIII.  Indicated  Horse  Power.  —  The  indicated  horse  power  should 
be  determined  from  the  average  mean  effective  pressure  of  diagrams 
taken  at  intervals  of  twenty  minutes,  and  at  more  frequent  intervals 
if  the  nature  of  the  test  makes  this  necessary,  for  each  end  of  each 
cylinder.  With  variable  loads,  such  as  those  of  engines  driving  gener- 
ators for  electric  railroad  work,  and  of  rubber-grinding  and  rolling-mill 
engines,  the  diagrams  cannot  be  taken  too  often.  In  cases  like  the 
latter,  one  method  of  obtaining  suitable  averages  is  to  take  a  series  of 
diagrams  on  the  same  blank  card  without  unhooking  the  driving  cord, 
and  apply  the  pencil  at  successive  intervals  of  ten  seconds  until  two 
minutes'  time  or  more  has  elapsed,  thereby  obtaining  a  dozen  or  more 
indications  in  the  time  covered.  This  tends  to  insure  the  determina- 
tion of  a  fair  average  for  that  period.  In  taking  diagrams  for  variable 
loads,  as  indeed  for  any  load,  the  pencil  should  be  applied  long  enough 
to  cover  several  successive  revolutions,  so  that  the  variations  produced 
by  the  action  of  the  governor  may  be  properly  recorded.  To 
determine  whether  the  governor  is  subject  to  what  is  called  "  racing  " 
or  "  hunting,"  a  "  variation  diagram  "  should  be  obtained;  that  is, 
one  in  which  the  pencil  is  applied  a  sufficient  time  to  cover  a  complete 
cycle  of  variations.  When  the  governor  is  found  to  be  working  in  this 
manner,  the  defect  should  be  remedied  before  proceeding  with  the  test. 


854  STEAM  POWER  PLANT  ENGINEERING 

It  is  seldom  necessary,  as  far  as  average  power  measurements  are 
concerned,  to  obtain  diagrams  at  precisely  the  same  instant  at  the 
two  ends  of  the  cylinder,  or  at  the  same  instant  on  all  the  cylinders, 
when  there  are  more  than  one.  All  that  is  required  is  to  take  the 
diagrams  at  regular  intervals.  Should  the  diagrams  vary  so  much 
among  themselves  that  the  average  may  not  be  a  fair  one,  it  signifies 
that  they  should  be  taken  more  frequently,  and  not  that  special  care 
should  be  employed  to  obtain  the  diagrams  of  each  set  at  precisely 
the  same  time.  When  diagrams  are  taken  during  the  time  when  the 
engine  is  working  up  to  speed  at  the  start,  or  when  a  study  of  valve 
setting  and  steam  distribution  is  being  made,  they  should  be  taken  at 
as  nearly  the  same  time  as  practicable.  In  cases  where  the  diagrams 
are  to  be  taken  simultaneously,  the  best  plan  is  to  have  an  operator 
stationed  at  each  indicator.  This  is  desirable,  even  where  an  electric 
or  other  device  is  employed  to  operate  all  the  instruments  at  once  ; 
for  unless  there  are  enough  operators,  it  is  necessary  to  open  the  indi- 
cator cocks  some  time  before  taking  the  diagrams  and  run  the  risk  of 
clogging  the  pistons  and  heating  the  high-pressure  springs  above  the 
ordinary  working  temperature. 

The  most  satisfactory  driving  rig  for  indicating  seems  to  be  some 
form  of  well-made  pantagraph,  with  driving  cord  of  fine  annealed  wire 
leading  to  the  indicator.  The  reducing  motion,  whatever  it  may  be, 
and  the  connections  to  the  indicator,  should  be  so  perfect  as  to  pro- 
duce diagrams  of  equal  lengths  when  the  same  indicator  is  attached  to 
either  end  of  the  cylinder,  and  produce  a  proportionate  reduction  of  the 
motion  of  the  piston  at  every  point  of  the  stroke,  as  proved  by  test. 

The  use  of  a  three-way  cock  and  a  single  indicator  connected  to  the 
two  ends  of  the  cylinder  is  not  advised,  except  in  cases  where  it  is 
impracticable  to  use  an  indicator  close  to  each  end.  If  a  three-way 
cock  is  used,  the  error  produced  should  be  determined  and  allowed  for. 

To  determine  the  average  power  developed  in  cases  where  the  engine 
starts  from  rest  during  the  progress  of  the  trial,  as  in  a  commercial 
test  of  a  plant  where  the  engine  runs  only  a  portion  of  the  twenty- 
four  hours,  a  number  of  diagrams  should  be  taken  during  the  period  of 
getting  up  speed  and  applying  the  working  load,  the  corresponding 
speed  for  each  set  of  diagrams  being  counted.  The  power  shown  by 
these  diagrams  for  the  proportionate  time  should  be  included  in  the 
average  for  the  whole  run,  and  the  duration  should  be  the  time  the 
throttle  valve  is  open. 

XIV.  Testing  Indicator  Springs.  —  To  make  a  perfectly  satisfactory 
comparison  of  indicator  springs  with  standards,  the  calibration  should 
be  made,  if  this  were  practical,  under  the  same  conditions  as  those 


APPENDIX  C  855 

pertaining  to  their  ordinary  use.  Owing  to  the  fact  that  the  pressure 
of  the  steam  in  the  indicator  cylinder  and  the  corresponding  temperature 
are  undergoing  continual  changes,  it  becomes  almost  impossible  to 
compare  the  springs  with  any  standard  under  such  conditions.  There 
must  be  a  constant  pressure  during  the  time  that  the  comparison  is 
being  made.  Although  the  best  that  can  be  done  is  not  altogether 
satisfactory,  it  seems  that  we  must  be  content  with  it.  To  bring  the 
conditions  as  nearly  as  possible  to  those  of  the  working  indicator,  the 
steam  should  be  admitted  to  the  indicator  as  short  a  time  as  practicable 
for  each  of  the  pressures  tried,  and  then  the  indicator  cock  should  be 
closed  and  the  steam  exhausted  therefrom  before  another  pressure  is 
tried.  By  this  means  the  parts  are  heated  and  cooled  somewhat  the 
same  as  under  the  working  conditions.  We  recommend,  therefore,  that 
for  each  required  pressure  the  first  step  be  to  open  and  close  the  indicator 
cock  a  number  of  times  in  quick  succession,  then  to  quickly  draw  the 
line  on  the  paper  for  the  desired  record,  observing  the  gauge  or  other 
standard  at  the  instant  when  the  line  is  drawn.  A  corresponding 
atmospheric  line  is  taken  immediately  after  obtaining  the  line  at  the 
given  pressure,  so  as  to  eliminate  any  difference  in  the  temperature  of 
the  parts  of  the  indicator.  This  appears  to  be  a  better  method  (although 
less  readily  carried  on  and  requiring  more  care)  than  the  one  heretofore 
more  commonly  used,  where  the  indicator  cock  is  kept  continually  open 
and  the  pressure  is  gradually  rising  or  falling  through  the  range  of 
comparison. 

The  calibration  should  be  made  for  at  least  five  points,  two  of  these 
being  for  the  pressure  corresponding  as  near  as  may  be  to  the  initial 
and  back  pressures,  and  three  for  intermediate  points  equally  distant. 

For  pressures  above  the  atmosphere,  the  proper  standard  recom- 
mended is  the  dead- weight  testing  apparatus,  or  a  reliable  mercury 
column,  or  an  accurate  steam  gauge  proved  correct,  or  of  known  error, 
by  either  of  these  standards.  For  pressures  below  the  atmosphere  the 
best  standard  to  use  is  a  mercury  column. 

The  correct  scale  of  spring  to  be  used  for  working  out  the  mean 
effective  pressure  of  the  diagrams  should  be  the  average  based  on  the 
calibration,  and  this  may  be  ascertained  in  the  manner  pointed  out 
below. 

XV.  Brake  Horse  Power.  —  This  term  applies  to  the  power  delivered 
from  the  fly-wheel  shaft  of  the  engine.  It  is  the  power  absorbed  by  a 
friction  brake  applied  to  the  rim  of  the  wheel  or  to  the  shaft.  A  form 
of  brake  is  preferred  that  is  self-adjusting  to  a  certain  extent,  so  that  it 
will  of  itself  tend  to  maintain  a  constant  resistance  at  the  rim  of  the 
wheel.  One  of  the  simplest  brakes  for  comparatively  small  engines, 


856 


STEAM  POWER  PLANT  ENGINEERING 


which  may  be  made  to  embody  this  principle,  consists  of  a  cotton  or 
hemp  rope,  or  a  number  of  ropes,  encircling  the  wheel,  arranged  with 
weighing  scales  or  other  means  for  showing  the  strain.  An  ordinary 
band  brake  may  also  be  constructed  so  as  to  embody  the  principle. 
The  wheel  should  be  provided  with  interior  flanges  for  holding  water 
used  for  keeping  the  rim  cool. 


FIG.  454. 


Rope  Brakes. 


FIG.  455. 


A  self-adjusting  rope  brake  is  illustrated  in  Fig.  454,  where  it  will 
be  seen  that,  if  the  friction  at  the  rim  of  the  wheel  increases,  it  will 
lift  the  weight  A,  which  action  will  diminish  the  tension  in  the  end  B 
of  the  rope  and  thus  prevent  a  further  increase  in  the  friction.  The 
same  device  can  be  used  for  a  band  brake  of  the  ordinary  construction. 
Where  space  below  the  wheel  is  limited,  a  cross  bar,  C,  supported  by 
a  chain  tackle  exactly  at  its  center  point  may  be  used  as  shown  in  Fig. 
455,  thereby  causing  the  action  of  the  weight  on  the  brake  to  be  up- 
ward. A  safety  stop  should  be  used  with  either  form,  to  prevent  the 
weights  being  accidentally  raised  more  than  a  certain  amount. 

The  water-friction  brake  is  specially  adapted  for  high  speeds  and  has 
the  advantage  of  being  self-cooling.  The  Alden  brake  is  also  self- 
cooling  and  is  capable  of  fine  adjustment. 

A  water-friction  brake  is  shown  in  Fig.  456.  It  consists  of  two  cir- 
cular disks,  A  and  B,  attached  to  the  shaft  C,  and  revolving  in  a  case, 
E,  between  fixed  planes.  The  space  between  the  disks  and  planes  is 


APPENDIX  C 


857 


supplied  with  running  water,  which  enters  at  D  and  escapes  at  the 
cocks  F,  G,  and  H.  The  friction  of  the  water  against  the  surfaces  con- 
stitutes a  resistance  which  absorbs  the  desired  power,  and  the  heat 
generated  within  is  carried  away  by  the  water  itself.  The  water  is 
thrown  outward  by  centrifugal  action  and  fills  the  outer  portion  of 
the  case.  The  greater  the  depth  of  the  ring  of  water  the  greater  the 


FIG.  456.    Alden  Absorption  Dynamometer. 

amount  of   power  absorbed.       By  suitably  adjusting  the  amount  of 
water  entering    and    leaving    any    desired    power   can   be    obtained. 
Water-friction  brakes  have  been  used  successfully  at  speeds  of  over 
20,000  revolutions  per  minute. 
For  description  of  the  Alden  brake,  see  Transactions,  Vol.  XI,  p.  958. 

XVI.  Quality  of  Steam.  —  When  ordinary  saturated  steam  is  used, 
its  quality  should  be  obtained  by  the  use  of  a  throttling  calorimeter 
attached  to  the  main  steam  pipe  near  the  throttle  valve.  When  the 
steam  is  superheated,  the  amount  of  superheating  should  be  found  by 
the  use  of  a  thermometer  placed  in  a  thermometer-well  filled  with  mer- 
cury, inserted  in  the  pipe.  The  sampling  pipe  for  the  calorimeter  should, 
if  possible,  be  attached  to  a  section  of  the  main  pipe  having  a  vertical 
direction,  with  the  steam  preferably  passing  upward,  and  the  sampling 
nozzle  should  be  made  of  a  half-inch  pipe,  having  at  least  20  one-eighth- 
inch  holes  in  its  perforated  surface.  The  readings  of  the  calorimeter 
should  be  corrected  for  radiation  of  the  instrument,  or  they  should  be 
referred  to  a  normal  reading,  as  pointed  out  below.  If  the  steam  is 


858  STEAM  POWER  PLANT  ENGINEERING 

superheated,  the  amount  of  superheating  should  be  obtained  by  refer- 
ring the  reading  of  the  thermometer  to  that  of  the  same  thermometer 
when  the  steam  within  the  pipe  is  saturated,  and  not  by  taking  the 
difference  between  the  reading  of  the  thermometer  and  the  temper- 
ature of  saturated  steam  at  the  observed  pressure  as  given  in  a  steam 
table. 

XVII.  Speed.  —  There  are  several  reliable  methods  of  ascertaining 
speed,  or  the  number  of  revolutions  of  the  engine   crank-shaft    per 
minute.     The  simplest  is  the  familiar  method  of  counting  the  number 
of  turns  for  a  period  of  one  minute  with  the  eye  fixed  on  the  second 
hand  of  a  timepiece.     Another  is  the  use  of  a  counter  held  for  a 
minute  or  a  number  of  minutes  against  the  end  of  the  main  shaft. 
Another  is  the  use  of  a  reliable  calibrated  tachometer  held  likewise 
against  the  end  of  the  shaft.     The  most  reliable  method,  and  the  one 
we  recommend,  is  the  use  of  a  continuous  recording  engine  register  or 
counter,  taking  the  total  reading  each  time  that  the  general  test  data 
are  recorded,  and  computing  the  revolutions  per  minute  corresponding 
to  the  difference  in  the  readings  of  the  instrument.     When  the  speed 
is  above  250  revolutions  per  minute,  it  is  almost  impossible  to  make  a 
satisfactory  counting  of  the  revolutions  without  the  use  of  some  form 
of  mechanical  counter. 

The  determination  of  variation  of  speed  during  a  single  revolution, 
or  the  effect  of  the  fluctuation  due  to  sudden  changes  of  the  load,  is 
also  desirable,  especially  in  engines  driving  electric  generators  used  for 
lighting  purposes.  There  is  at  present  no  recognized  standard  method 
of  making  such  determinations,  and  if  such  are  desired,  the  method 
employed  may  be  devised  by  the  person  making  the  test  and  described 
in  detail  in  the  report. 

XVIII.  Recording  the  Data.  —  Take  note  of  every  event  connected 
with  the  progress  of  the  trial  whether  it  seems  at  the  time  to  be 
important   or   unimportant.      Record   the   time   of   every   event   and 
time   of  taking  every   weight   and  every   observation.      Observe  the 
pressures,  temperatures,  water    heights,  speeds,  etc.,  every  twenty  or 
thirty  minutes  when  the  conditions  are  practically  uniform,   and  at 
much  more  frequent  intervals  if  the  conditions  vary.     Observations 
which   concern   the   feed-water   measurements   should   be   made   with 
special  care  at  the  expiration  of  each  hour  of  the  trial,  so  as  to  divide 
the  tests  into  hourly  periods  and  show  the  uniformity  of  the  conditions 
and  results  as  the  test  goes  forward.     Where  the  water  discharged  from 
a  surface  condenser  is  weighed  it  may  be  advisable  to  divide  the  test 
by  this  means  into  periods  of  less  than  one  hour. 


APPENDIX  C  859 

The  data  and  observations  of  the  test  should  be  kept  on  properly 
prepared  blanks  or  in  notebooks  containing  columns  suitably  arranged 
for  a  clear  record.  As  different  observers  have  their  own  individual 
ideas  as  to  how  such  records  should  be  kept,  no  special  form  of  log 
sheet  is  given  as  a  necessary  part  of  the  code. 

XIX.  Uniformity  of  Conditions.  —  In  a  test   having  for  an  object 
the    determination   of   the    maximum   economy   obtainable   from    an 
engine,  or  where  it  is  desired  to  ascertain  with  special  accuracy  the 
effect  of  predetermined  conditions  of  operation,  it  is  important  that  all 
the  conditions  under  which  the  engine  is  operated  should  be  main- 
tained uniformly  constant.     This  requirement  applies  especially  to  the 
pressure,  the  speed,  the  load,  the  rate  of  feeding  the  various  supplies  of 
water,  the  height  of  water  in  the  gauge  glasses,  and  the  depth  of  water 
in  the  feed-water  reservoir. 

XX.  Analysis  of  Indicator  Diagrams.  —  (a)  Steam  accounted  for  by 
the  Indicator :  The  simplest  method  of  computing  the  steam  accounted 
for  by  the  indicator  is  the  use  of  the  formula 

E)  XWc-(H  +  E)XWh], 

which  gives  the  weight  in  pounds  per  indicated  horse  power  per  hour. 
In  this  formula  the  symbol  "  M.E.P."  refers  to  the  mean  effective 
pressure.  In  multiple-expansion  engines  this  is  the  combined  mean 
effective  pressure  referred  to  the  cylinder  in  question.  The  symbol  C 
refers  to  the  proportion  of  the  stroke  completed  at  points  on  the 
expansion  line  of  the  diagram  near  the  actual  cut-off  or  release;  the 
symbol  H  to  the  proportion  of  compression;  and  the  symbol  E  to 
the  proportion  of  clearance;  all  of  which  are  determined  from  the  indi- 
cator diagram.  The  symbol  We  refers  to  the  weight  of  one  cubic  foot 
of  steam  at  the  cut-off  or  release  pressure;  and  the  symbol  Wh  to  the 
weight  of  one  cubic  foot  of  steam  at  the  compression  pressure;  these 
weights  being  taken  from  steam  tables  of  recognized  accuracy.  The 
points  near  the  cut-off  and  release  on  the  expansion  line  and  the  point 
on  the  compression  line  are  located  as  shown  on  the  sample  diagram, 
Fig.  457.  They  are  the  points  in  the  case  of  the  expansion  and  com- 
pression lines  of  the  diagram  which  mark  the  complete  closure  of  the 
valve.  The  point  near  the  cut-off,  for  example,  lies  where  the  curve 
of  expansion  begins  after  the  rounding  of  the  diagram  due  to  the  wire- 
drawing which  occurs  while  the  valve  is  closing.  This  cut-off  may  be 
located  by  finding  the  point  where  the  curve  is  tangent  to  a  hyper- 
bolic curve. 


860  STEAM  POWER  PLANT  ENGINEERING 

Should  the  point  in  the  compression  curve  be  at  the  same  height  as 
the  point  in  the  expansion  curve,  then  We  =  Wh,  and  the  formula 
becomes 


in  which  (C  —  H)  represents  the  distance  between  the  two  points 
divided  by  the  length  of  the  diagram. 

When  the  load  and  all  other  conditions  are  substantially  uniform,  it 
is  unnecessary  to  work  up  the  steam  accounted  for  by  the  indicator 


Release^ 

'Compressioa 
Atmospheric  Line 

FIG.  457.    Showing  Points  where  "  Steam  Accounted  for  by  Indicator  "  is 

Computed. 

from  all  the  diagrams  taken.  Five  or  more  sample  diagrams  may  be 
selected  and  the  computations  based  on  the  samples  instead  of  on  the 
whole. 

(6)  Sample  Indicator  Diagrams :  In  order  that  the  report  of  a  test 
may  afford  complete  information  regarding  the  conditions  of  the  test, 
sample  indicator  diagrams  should  be  selected  from  those  taken  and 
copies  appended  to  the  tables  of  results.  In  cases  where  the  engine  is 
of  the  multiple-expansion  type  these  sample  diagrams  may  also  be 
arranged  in  the  form  of  a  "  combined  "  diagram. 

(c)  The  Point  of  Cut-off:  The  term  "  cut-off  "  as  applied  to  steam 
engines,  although  somewhat  indefinite,  is  usually  considered  to  be  at  an 
earlier  point  in  the  stroke  than  the  beginning  of  the  real  expansion  line. 
That  the  cut-off  point  may  be  defined  in  exact  terms  for  commercial 
purposes,  as  used  in  steam-engine  specifications  and  contracts,  the 
Committee  recommends  that,  unless  otherwise  specified,  the  commercial 
cut-off,  which  seems  to  be  an  appropriate  expression  for  this  term,  be 
ascertained  as  follows:  Through  a  point  showing  the  maximum  pressure 


APPENDIX  C 


861 


during  admission  draw  a  line  parallel  to  the  atmospheric  line.  Through 
the  point  on  the  expansion  line  near  the  actual  cut-off,  referred  to  in 
Section  XX  (a),  draw  a  hyperbolic  curve.  The  point  where  these  two 
lines  intersect  is  to  be  considered  the  commercial  cut-off  point.  The 
percentage  is  then  found  by  dividing  the  length  of  the  diagram  measured 
to  this  point  by  the  total  length  of  the  diagram  and  multiplying  the 
result  by  100. 

E    C  B  A 


H   G  F 

FIG.  458.      Four  Valve  Engine,   Slow  Speed,  Commercial 


The  principle  involved  in  locating  the  commercial  cut-off  is  shown  in 
Figs.  458  and  459,  the  first  of  which  represents  a  diagram  from  a  slow- 
speed  Corliss  engine  and  the  second  a  diagram  from  a  single-valve 
high-speed  engine.  In  the  latter  case  where,  owing  to  the  fling  of  the 


H    G 
FIG.  459.     Single  Valve  Engine,  High-Speed,  Com- 


mercial Cut-off  = 


BC 
AC" 


pencil,  the  steam  line  vibrates,  the  maximum  pressure  is  found  by 
taking  a  mean  of  the  vibrations  at  the  highest  point. 

The  commercial  cut-off  as  thus  determined  is  situated  at  an  earlier 


862  STEAM  POWER   PLANT  ENGINEERING 

point  of  the  stroke  than  the  actual  cut-off  referred  to  in  computing  the 
"  steam  accounted  for  "  by  the  indicator  in  Section  XX  (a). 

(d)  Ratio  of  Expansion :  The  ratio  of  expansion  for  a  simple  engine 
is  determined  by  dividing  the  volume  corresponding  to  the  piston  dis- 
placement, including  clearance,  by  the  volume  of  the  steam  at  the  com- 
mercial cut-off,  including  clearance. 

In  a  multiple-expansion  engine  it  is  determined  by  dividing  the  net 
volume  of  the  steam  indicated  by  the  low-pressure  diagram  at  the  end 
of  the  expansion  line,  assumed  to  be  continued  to  the  end  of  the  stroke, 
by  the  net  volume  of  the  steam  at  the  maximum  pressure  during  admis- 
sion to  the  high-pressure  cylinder. 

(e)  Diagram  Factor :  The  diagram  factor  is  the  proportion  borne  by 
the  actual  mean  effective  pressure  measured  from  the  indicator  diagram 
to  that  of  a  diagram  in  which  the  various  operations  of  admission, 
expansion,  release,  and  compression  are  carried  on  under  assumed  con- 
ditions.    The  factor  recommended  refers  to  an  ideal  diagram  which 
represents  the  maximum  power  obtainable  from  the  steam  accounted 
for  by  the  indicator  diagrams  at  the  point  of  cut-off,  assuming  first  that 
the  engine  has  no  clearance;  second,  that  there  are  no  losses  through 
wire-drawing  the  steam  either  during  the  admission  or  the  release; 
third,  that  the  expansion  line  is  a  hyperbolic  curve;  and  fourth,  that  the 
initial  pressure  is  that  of  the  boiler  and  the  back  pressure  that  of  the 
atmosphere  for  a  non-condensing  engine  and  of  the  condenser  for  a 
condensing  engine. 

The  diagram  factor  is  useful  for  comparing  the  steam  distribution 
losses  in  different  engines,  and  is  of  special  use  to  the  engine  designer, 
for  by  multiplying  the  mean  effective  pressure  obtained  from  the 
assumed  theoretical  diagrams  by  it  he  will  obtain  the  actual  mean 
effective  pressure  that  should  be  developed  in  an  engine  of  the  type 
considered.  The  expansion  and  compression  curves  are  taken  as 
hyperbolas,  because  such  curves  are  ordinarily  used  by  engine  builders 
in  their  work,  and  a  diagram  based  on  such  curves  will  be  more  useful  to 
them  than  one  where  the  curves  are  constructed  according  to  a  more 
exact  law. 

In  cases  where  there  is  a  considerable  loss  of  pressure  between  the 
boiler  and  the  engine,  as  where  steam  is  transmitted  from  a  central 
plant  to  a  number  of  consumers,  the  pressure  of  the  steam  in  the  supply 
main  should  be  used  in  place  of  the  boiler  pressure  in  constructing  the 
diagrams. 

XXI.  Standards  of  Economy  and  Efficiency.  —  The  hourly  consump- 
tion of  heat,  determined  by  employing  the  actual  temperature  of  the 
feed  water  to  the  boiler,  as  pointed  out  in  Article  IX  of  the  Code,  divided 


APPENDIX  C  863 

by  the  indicated  and  brake  horse  power,  that  is,  the  number  of  heat 
units  consumed  per  indicated  and  per  brake  horse  power  per  hour,  is 
the  standard  of  engine  efficiency  recommended  by  the  Committee. 
The  consumption  per  hour  is  chosen  rather  than  the  consumption  per 
minute,  so  as  to  conform  with  the  designation  of  time  applied  to  the 
more  familiar  units  of  coal  and  water  measurement  which  have  hereto- 
fore been  used.  The  British  standard,  where  the  temperature  of  the 
feed  water  is  taken  as  that  corresponding  to  the  temperature  of  the 
back-pressure  steam,  allowance  being  made  for  any  drips  from  jackets 
or  reheaters,  is  also  included  in  the  tables. 

It  is  useful  in  this  connection  to  express  the  efficiency  in  its  more 
scientific  form,  or  what  is  called  the  "  thermal  efficiency  ratio."  The 
thermal  efficiency  ratio  is  the  proportion  which  the  heat  equivalent  of 
the  power  developed  bears  to  the  total  amount  of  heat  actually  con- 
sumed, as  determined  by  test.  The  heat  converted  into  work  repre- 
sented by  one  horse  power  is  1,980,000  foot-pounds  per  hour,  and  this 
divided  by  778  equals  2545  British  thermal  units.  Consequently  the 
thermal  efficiency  ratio  is  expressed  by  the  fraction 

2545 


2545 

B.T.U.  per  H.P.  per  hour' 


XXII.  Heat  Analysis.  —  For  certain  scientific  investigations  it  is 
useful  to  make  a  heat  analysis  of  the  diagram  to  show  the  interchange 
of  heat  from  steam  to  cylinder  walls,  etc.,  which  is  going  on  within  the 
cylinder.     This  is  unnecessary. for  commercial  tests. 

XXIII.  Temperature-Entropy   Diagram.  —  The  study   of  the   heat 
analysis  is  facilitated  by  the  use  of  the  temperature-entropy  diagram 
in  which   areas  represent   quantities   of  heat,  the   coordinates  being 
the  absolute  temperature  and  entropy.     Such  a  diagram  is  shown  in 
Fig.  460. 

When  the  quantities  given  in  the  steam  tables  are  plotted,  two 
curves,  AA  and  BB,  are  obtained  which  may  be  termed  the  water  line 
and  the  steam  line,  AA  being  the  logarithmic  curve  if  the  specific  heat 
of  the  water  is  taken  as  constant.  The  diagram  refers  to  a  unit  weight 
of  the  agent,  and  the  heat  necessary  to  raise  a  pound  of  water  from 
the  temperature  ma  to  the  temperature  pa'  and  evaporate  it  at 
that  temperature  is  represented  by  the  area  aa'b'qm.  If  the  steam 
be  now  expanded  adiabatically  the  temperature  will  fall  to  qs  and 

x  per  cent  =  —  will  remain  as  steam,  the  rest  being  liquefied.     If  the 
ab 

steam  is  now  rejected,  it  carries  away  with  it  the  heat  sqma,  the  work 


864 


STEAM  POWER   PLANT  ENGINEERING 


area  being  afb'sa,  from  which  must  be  deducted  the  work  w  (ex- 
pressed in  heat  units)  to  pump  a  pound  of  water  into  the  boiler.  The 
efficiency  of  this  cycle  is  evidently 


h 


in  which 


x  = 


ar 


cfV 


ab 


By  the  action  of  the  walls  a  portion  of  the  steam  is  liquefied  prior 
to  the  expansion,  which  therefore  begins  at  e,  and  since  the   cooling 


q      n 


FIG.  460.  FIG.  461. 

Temperature-Entropy  Diagrams. 

action  of  the  walls  continues,  the  expansion  line  falls  off  to  ef,  from 
which  point  a.  reverse  action  takes  place  and  the  expansion  line  bends 


APPENDIX  C  865 

over  to  g.  Finally,  since  the  release  takes  place  before  the  condenser 
temperature  is  reached,  the  heat  rejection  starts  at  g,  following  a  line 
of  equal  volume  until  the  exhaust  port  temperature  is  reached  at  /. 
If  heat  is  added  during  expansion  enough  to  keep  the  steam  theo- 
retically saturated,  as,  for  example,  by  a  water  jacket,  such  additional 
heat  is  represented  by  the  area  b'bnq,  and  the  additional  work  obtained 
by  the  triangle  b'bs.  If  the  steam  is  superheated  sufficiently  to  give 
by  expansion  theoretically  dry  steam  at  the  end,  such  additional  heat 
is  represented  by  the  area  b'vnq  and  the  additional  work  by  b'vbs. 
Neither  of  these  extra  amounts  of  work  is  realized  in  practice,  and 
it  is  evident  from  the  diagram  that  the  heat  thus  applied  is  in  both 
cases  less  efficient  than  in  the  principal  cycle.  Nevertheless  the  action 
in  each  case  is  to  bring  the  point  e  nearer  the  point  b'  and  to  effect  a 
notable  net  economy. 

The  Carnot  cycle  would  be  obtained  if  in  the  Rankine  cycle  the 
rejection  of  heat  were  stopped  at  r  and  the  temperature  of  the  mix- 
ture raised  to  a'  by  compression.  This  cannot  be  practically  accom- 
plished, but  a  system  of  feed-water  heaters  has  been  suggested  and 
exemplified  in  the  Nordberg  engine,  which  is  theoretically  a  close 
equivalent  to  it.  Where  steam  is  expanded  in  say  three  cylinders,  the 
feed  water  may  be  successively  heated  from  the  receiver  intermediate 
between  each  pair,  the  effect  of  which  is  illustrated  in  Fig.  461.  The 
expansion  line  follows  the  heavy  line,  being  carried  over  to  y  by  the 
first  feed-water  heater  and  to  yf  by  the  second  feed-water  heater.  With 
an  infinite  number  of  such  feed-water  heaters,  the  line  yy'  would  be 
parallel  to  aar,  and  the  cycle  equivalent  to  that  of  Carnot. 

XXIV.  Ratio  of  Economy  of  an  Engine  to  that  of  an  Ideal  Engine.  — 
The  ideal  engine  recommended  for  obtaining  this  ratio  is  that  which 
was  adopted  by  the  Committee  appointed  by  the  Civil  Engineers  of 
London  to  consider  and  report  a  standard  thermal  efficiency  for 
steam  engines.  This  engine  is  one  which  follows  the  Rankine  cycle, 
where  steam  at  a  constant  pressure  is  admitted  into  the  cylinder 
with  no  clearance,  and  after  the  point  of  cut-off  is  expanded  adiabat- 
ically  to  the  back  pressure.  In  obtaining  the  economy  of  this  engine 
the  feed  water  is  assumed  to  be  returned  to  the  boiler  at  the  exhaust 
temperature.  Such  a  cycle  is  preferable  to  the  Carnot  for  the  purpose 
at  hand,  because  the  Carnot  cycle  is  theoretically  impossible  for  an 
engine  using  superheated  steam  produced  at  a  constant  pressure,  and 
the  gain  in  efficiency  for  superheated  steam  corresponding  to  the 
Carnot  efficiency  will  be  much  greater  than  that  possible  for  the  actual 
cycle. 

The  ratio  of  the  economy  of  an  engine  to  that  of  the  ideal 
engine  is  obtained  by  dividing  the  heat  consumption  per  indicated 


866  STEAM  POWER  PLANT  ENGINEERING 

horse  power  per  minute  for  the  ideal  engine  by  that  of  the  actual 
engine. 

XXV.  Miscellaneous.  —  In  the  case  of  tests  of  combined  engines 
and  boiler  plants,  where  the  full  data  of  the  boiler  performance  is  to 
be  determined,  reference  should  be  made  to  the  directions  given  by  the 
Boiler  Test  Committee  of  the  Society,  Code  of  1899.     (See  Vol.  XXI, 
p.  34.) 

In  tests  made  for  scientific  research,  and  in  those  made  on  special 
forms  of  engines,  the  line  of  procedure  must  be  varied  according  to  the 
special  objects  in  view,  and  it  has  been  deemed  unnecessary  to  go  into 
particulars  applying  to  such  tests. 

In  testing  steam  pumping  engines  and  locomotives  in  accordance 
with  the  standard  methods  of  conducting  such  tests,  recommended  by 
the  committees  of  the  Society,  reference  should  be  made  to  the  reports 
of  those  committees  in  the  Transactions,  Vol.  XII,  p.  530,  and  in  Vol. 
XIV,  p.  1312. 

XXVI.  Report  of  Test.  —  The  data  and  results  of  the  test  should  be 
reported  in  the  manner  and  in  the  order  outlined  in  one  of  the  fol- 
lowing tables,  the  first  of  which  gives,  it  is  hoped,  a  complete  sum- 
mary of  all  the  data  and  results  as  applied  not  only  to  the  standard 
heat-unit  test  but  also  to  tests  of  combined  engine  and  boiler  for  deter- 
mining all  questions  of  performance,  whatever  the  class  of  service; 
the  second  refers  to  a  short  form  of  report  giving  the  necessary  data 
and  results  for  the  standard  heat  test;    and  the  third  to  a  short  form 
of  report  for  a  feed-water  test.     It  is  the  intention  that  the  tables 
should  be  full  enough  to  apply  to  any  type  of  engine,  but  where  not 
so,  or  where  special  data  and  results  are  determined,  additional  results 
may   be  inserted   under  the   appropriate   headings.     Although   these 
forms  are  arranged  so  as  to  be  used  for  expressing  the  principal  data 
and  results  of  tests  of  pumping  engines  and  locomotives,  as  well  as 
for  all  other  classes  of  steam  engines,  it  is  not  the  intention  that  they 
shall  supplant  the  forms  recommended  by  the  committees  on  Duty 
Trials  and  Locomotives  in  cases  where  the  full  report  of  a  test  of  such 
engines  is  desired. 

It  is  recommended  that  any  report  be  supplemented  by  a  chart  in 
which  the  data  of  the  test  are  graphically  presented.  (As  an  example 
of  such  a  chart  as  applied  to  a  boiler  test,  see  Vol.  XXI,  p.  104.) 

TABLE  NO.  1. 
Not  reprinted  here.     See  Trans.  A.S.M.E.  24-702. 


APPENDIX  C  867 

TABLE  NO.  2. 
DATA  AND  RESULTS  OF  STANDARD  HEAT  TEST  OF  STEAM  ENGINE. 

Arranged  according  to  the  Short  Form  advised  by  the  Engine  Test  Committee  of  the 
American  Society  of  Mechanical  Engineers.     Code  of  1902. 

1.  Made  by of 

on  engine  located  at 

to  determine , 

2.  Date  of  trial 

3.  Type  and  class  of  engine ;  also  of  condenser 


1st  Cyl.         2d  Cyl.         3d  Cyl. 

4.  Dimensions  of  main  engine 

(a)  Diameter  of  cylinder in. 

(6)  Stroke  of  piston ft. 

(c)  Diameter  of  piston  rod in. 

(d)  Average  clearance p.  c. 

(e)  Ratio    of    volume    of    cylinder    to    high- 

pressure  cylinder 

(/)  Horse-power  constant  for  one  pound  mean 
effective  pressure  and  one  revolution 
per  minute 

5.  Dimensions  and  type  of  auxiliaries 


Total  Quantities,  Time,  etc. 

6.  Duration  of  test hours. 

7.  Total  water  fed  to  boilers  from  main  source  of  supply pounds. 

8.  Total  water  fed  from  auxiliary  supplies: 

(a) 

(&) 

(c) 

9.  Total  water  fed  to  boilers  from  all  sources pounds. 

10.  Moisture  in  steam  or  superheating  near  throttle p.  c.  or  deg. 

11.  Factor  of  correction  for  quality  of  steam 

12.  Total  dry  steam  consumed  for  all  purposes pounds 

Hourly  Quantities. 

13.  Water  fed  from  main  source  of  supply " 

14.  Water  fed  from  auxiliary  supplies: 

(a) 

(6) • 

(c) 

15.  Total  water  fed  to  boilers  per  hour " 

16.  Total  dry  steam  consumed  per  hour " 

17.  Loss  of  steam  and  water  per  hour  due  to  drips  from  main  steam 

pipes  and  to  leakage  of  plant " 

18.  Net  dry  steam  consumed  per  hour  by  engine  and  auxiliaries " 


868  STEAM  POWER  PLANT  ENGINEERING 

Pressures  and  Temperatures  (Corrected'). 

19.  Pressure  in  steam  pipe  near  throttle  by  gauge  ..................  Ib.  per  sq.  in. 

20.  Barometric  pressure  of  atmosphere  in  inches  of  mercury  .........  inches. 

21.  Pressure  in  receivers  by  gauge  ................................  Ib.  per  sq.  in. 

22.  Vacuum  in  condenser  in  inches  of  mercury  .....................  inches. 

23.  Pressure  in  jackets  and  reheaters  by  gauge  .....................  Ib.  per  sq.  in. 

24.  Temperature  of  main  supply  of  feed  water  .....................       degrees  F. 

25.  Temperature  of  auxiliary  supplies  of  feed  water: 


(c)  ...................................................... 

26.  Ideal  feed-water  temperature  corresponding  to  pressure  of  steam 

in  the  exhaust  pipe,  allowance  being  made  for  heat  derived 

from  jacket  or  reheater  drips  ............................  " 

Data  Relating  to  Heat  Measurement. 

27.  Heat  units  per  pound  of  feed  water,  main  supply  ...............  B.T.IT. 

28.  Heat  units  per  pound  of  feed  water,  auxiliary  supplies: 

(a)  ..................................................... 

(6)  ...................................................... 

(c)  ...................................................... 

29.  Heat  units  consumed  per  hour,  main  supply  ...................  " 

30.  Heat  units  consumed  per  hour,  auxiliary  supplies: 

(a)  ..................................................... 

(&)  ...................................................... 

(0  ..................  ..................  .................. 

31.  Total  heat  units  consumed  per  hour  for  all  purposes  ............  " 

32.  Loss  of  heat  per  hour  due  to  leakage  of  plant,  drips,  etc  .........  " 

33.  Net  heat  units  consumed  per  hour: 

(a)  By  engine  alone  ......................................  " 

(6)  By  auxiliaries  ........................................  " 

34.  Heat  units  consumed  per  hour  by  engine  alone,  reckoned  from 

temperature  given  in  line  26  ............................  " 

Indicator  Diagrams. 

1st  Cyl.         2d  Cyl.         3d  Cyl. 

35.  Commercial  cut-off  in  per  cent  of  stroke  ........ 

36.  Initial  pressure  in  pounds  per  square  inch  above 

atmosphere  ............................. 

37.  Back  pressure  at  mid-stroke  above  or  below  at- 

mosphere in  pounds  per  square  inch  ....... 

38.  Mean  effective  pressure  in  pounds  per  square 

inch  .................................... 

39.  Equivalent   mean  effective   pressure   in   pounds 

per  square  inch  : 
(a)  Referred  to  first  cylinder  ................ 

(6)  Referred  to  second  cylinder  ............. 

(c)  Referred  to  third  cylinder  ............... 


APPENDIX  C  869 

IstCyl.         2dCyI.         3d  Cyl. 

40.  Pressures   and   percentages   used   in   computing 

the  steam  accounted  for  by  the  indicator 
diagrams,  measured  to  points  on  the  expan- 
sion and  compression  curves 

Pressure  above  zero  in  pounds  per  square  inch: 

(a)  Near  cut-off 

(6)  Near  release 

(c)  Near  beginning  of  compression 

Percentage  of  stroke  at  points  where  pressures 
are  measured: 

(a)  Near  cut-off 

(6)  Near  release 

(c)  Near  beginning  of  compression 

41.  Steam  accounted  for  by  indicator  in  pounds  per 

I.H.P.  per  hour: 

(a)  Near  cut-off . . 

(6)  Near  release 

42.  Ratio  of  expansion 

Speed 

43.  Revolutions  per  minute revolutions. 

Power. 

44.  Indicated  horse  power  developed  by  main-engine  cylinders: 

First  cylinder horse  power. 

Second  cylinder " 

Third  cylinder " 

Total " 

45.  Brake  horse  power  developed  by  engine " 

Standard  Efficiency  and  other  Results.* 

46.  Heat  units  consumed  by  engine  and  auxiliaries  per  hour: 

(a)  per  indicated  horse  power B.T.U. 

(6)  per  brake  horse  power " 

47.  Equivalent  standard  coal  in  pounds  per  hour: 

(a)  per  indicated  horse  power pounds. 

(6)  per  brake  horse  power " 

48.  Heat  units  consumed  by  main  engine  per  hour  corresponding  to 

ideal  maximum  temperature  of  feed  water  given  in  line  26, 

British  standard: 

(a)  per  indicated  horse  power B.T.U. 

(6)  per  brake  horse  power " 

49.  Dry  steam  consumed  per  indicated  horse  power  per  hour: 

(a)  Main  cylinders,  including  jackets pounds. 

(6)  Auxiliary  cylinders " 

(c)  Engine  and  auxiliaries " 

*  The  horse  power  referred  to  above  (items  46-50)  is  that  of  the  main  engine, 
exclusive  of  auxiliaries. 


870  STEAM  POWER  PLANT  ENGINEERING 

50.  Dry  steam  consumed  per  brake  horse  power  per  hour: 

(a)  Main  cylinders,  including  jackets pounds. 

(6)  Auxiliary  cylinders 

(c)  Engine  and  auxiliaries " 

51.  Percentage  of  steam  used  by  main-engine  cylinders  accounted 

for  by  indicator  diagrams,  near  cut-off  of  high-pressure  cylinder         per  cent. 

Additional  Data. 

Add  any  additional  data  bearing  on  the  particular  objects  of  the  test  or  relating 
to  the  special  class  of  service  for  which  the  engine  is  used.  Also  give  copies  of 
indicator  diagrams  nearest  the  mean,  and  the  corresponding  scales. 


TABLE  NO.  3. 

DATA  AND  RESULTS  OF  FEED-WATER  TEST  OF  STEAM  ENGINE. 

Arranged  according  to  the  Short  Form  advised  by  the  Engine  Test  Committee  of 
the  American  Society  of  Mechanical  Engineers.     Code  of  1902. 


1.  Made  by of 

on  engine  located  at 

to  determine 


2.  Date  of  trial 

3.  Type  of  engine  (simple,  compound,  or  other  multiple  expansion;  condensing 

or  non-condensing) 

4.  Class  of  engine  (mill,  marine,  locomotive,  pumping,  electric,  or  other) 

5.  Rated  power  of  engine 

6.  Name  of  builders 

7.  Number  and  arrangement  of  cylinders  of  engine;  how  lagged;  type  of  valvea 

and  of  condensers 

8.  Dimensions  of  engine lst  Cyl'    2d  Cyl'    3d  CyL 

(a)  Single  or  double  acting 

(6)  Cylinder  dimensions: 

Bore in. 

Stroke ft. 

Diameter  of  piston  rod in. 

Diameter  of  tail  rod in. 

(c)  Clearance  in  per  cent  of  volume  displaced 

by  piston  per  stroke: 

Head  end 

Crank  end 

Average 

(d)  Ratio    of    volume    of    each    cylinder    to 

volume  of  high-pressure  cylinder 

(e)  Horse-power  constant  for  one  pound  mean 

effective  pressure  and  one  revolution 
per  minute 


APPENDIX  C  871 

Total  Quantities,  Time,  etc. 
9.  Duration  of  test hours. 

10.  Water  fed  to  boilers  from  main  source  of  supply pounds. 

11.  Water  fed  from  auxiliary  supplies: 

(a) 

(6).. 

(c) 

12.  Total  water  fed  from  all  sources 

13.  Moisture  in  steam  or  superheating  near  throttle  * p.  c.,  or  deg. 

14.  Factor  of  correction  for  quality  of  steam 

15.  Total  dry  steam  consumed  for  all  purposes pounds. 

Hourly  Quantities. 

16.  Water  fed  from  main  source  of  supply 

17.  Water  fed  from  auxiliary  supplies: 

(a) 

(6) 

(C) : 

18.  Total  water  fed  to  boilers  per  hour 

19.  Total  dry  steam  consumed  per  hour " 

20.  Loss  of  steam  and  water  per  hour  due  to  leakage  of  plant,  drips, 

etc 

21.  Net  dry  steam  consumed  per  hour  by  engine  and  auxiliaries " 

22.  Dry  steam  consumed  per  hour: 

(a)  Main  cylinders 

(6)  Jackets  and  reheaters " 

Pressures  and  Temperatures  (Corrected'). 

23.  Steam  pipe  pressure  near  throttle,  by  gauge Ib.  per  sq.  in. 

24.  Barometric  pressure  of  atmosphere  in  inches  of  mercury inches. 

25.  Pressure  in  first  receiver  by  gauge Ib.  per  sq.  in. 

26.  Pressure  in  second  receiver  by  gauge 

27.  Vacuum  in  condenser: 

(a)  In  inches  of  mercury inches. 

(6)  Corresponding  total  pressure Ib.  per  sq.  in. 

28.  Pressure  in  steam  jackets  by  gauge Ib.  per  sq.  in. 

29.  Pressure  in  reheater  by  gauge 

30.  Superheating  of  steam  in  first  receiver degrees  F. 

31.  Superheating  of  steam  in  second  receiver 

Indicator  Diagrams. 

1st  Cyl.     2d  Cyl.     3d  Cyl. 

32.  Commercial  cut-off  in  per  cent  of  stroke 

33.  Initial  pressure  in  pounds  per  square  inch  above 

atmosphere 

*  In  case  of  superheated  steam  engines,  determine,  if  practicable,  the  temperature 
of  the  steam  in  each  cylinder. 


872  STEAM  POWER  PLANT  ENGINEERING 

IstCyl.        2dCyl.         3dCyl. 

34.  Back  pressure  at    mid-stroke    above   or    below 

atmosphere  in  pounds  per  square  inch  .... 

35.  Mean  effective  pressure  in  pounds  per  square 

inch 

36.  Equivalent  mean  effective  pressure  in  pounds  per 

square  inch  per  indicated  horse  power 

(a)  Referred  to  first  cylinder. 
(6)  Referred  to  second  cylinder, 
(c)  Referred  to  third  cylinder. 

37.  Pressures   and   percentages   used   in   computing 

the  steam  accounted  for  by  the  indicator 
diagrams,  measured  to  points  on  the  expan- 
sion and  compression  curves 

Pressures  above  zero  in  pounds  per  square  inch: 

(a)  Near  cut-off 

(6)  Near  release 

(c)  Near  beginning  of  compression 

Percentage  of  stroke  at  points  where  pressures 
are  measured: 

(a)  Near  cut-off , 

(6)  Near  release 

(c)  Near  beginning  of  compression 

38.  Aggregate   M.E.P.   in   pounds   per  square   inch 

referred  to  each  cylinder  given  in  heading 

39.  Mean    back  pressure    above    zero,  pounds    per 

square  inch 

40.  Steam  accounted  for  in  pounds  per  indicated 

horse  power  per  hour: 

(a)  Near  cut-off 

(6)  Near  release 

41.  Ratio  of  expansion: 

(a)  Commercial 

(6)  Ideal 


Speed. 

42.  Revolutions  per  minute revolutions. 

43.  Piston  speed  per  minute feet. 

Power. 

44.  Indicated  horse  power  developed  by  main-engine  cylinders: 

First  cylinder horse  power. 

Second  cylinder 

Third  cylinder " 


Total 


APPENDIX  C  873 

Efficiency  Results. 

45.  Dry  steam  consumed  per  indicated  horse  power  per  hour: 

(a)  Main  cylinder,  including  jackets pounds. 

(6)  Auxiliary  cylinders,  etc 

(c)  Engine  and  auxiliaries 

46.  Percentage  of  steam  used  by  main-engine  cylinders  accounted  for 

by  indicator  diagrams : 

IstCyl.    2dCyl.    SdCyl. 

(a)  Near  cut-off 

(6)  Near  release 

Sample  Diagrams. 

Copies  of  indicator  diagrams,  nearest  the  mean,  with  corresponding  scales,  should 
be  given  in  connection  with  table. 


874 


STEAM  POWER  PLANT  ENGINEERING 


Density 
Weight  per 
Cubic  Foot, 
Pounds. 

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APPENDIX  E. 

EQUIVALENT  VALUES  OF  ELECTRICAL  AND  MECHANICAL  UNITS. 


1  KILOWATT  HOUR  = 
1,000  watt  hours 
1.34  horse-power  hours 
2,654,200  foot-pounds 
3,600,000  joules 
367,000  kilogram  meters 

3.53  pounds  of  water  evaporated 
from  and  at  212°  F. 

1  KILOWATT  = 
1,000  watts 

1.34  horse  power 
1.358  cheval-vapeur 
2,654,200  foot-pounds  per  hour 
44,240  foot-pounds  per  minute 
737.3  foot-pounds  per  second 
3,412  B.T.U.  per  hour 

56.9  B.T.U.  per  minute 
0.948  B.T.U.  per  second 
3.53  pounds  of  water  evaporated 
from  and  at  212°  F. 

1  JOULE  = 

1  watt  second 

0.000000278  kilowatt  hour 
0.102  kilogram  meter 

0.0009477      B.T.U. 
0.7373  foot-pound 

1  WATT  = 

1  joule  per  second 

0.00134  horse  power 
3.412      B.T.U.  per  hour 
0.7373    foot-pound  per  second 
0.0035    pound  water  evaporated  from 
and  at  212°  F. 

1  KILOGRAM-METER  = 
7.233  foot-pounds 

0.00000365  horse-power  hour 
0.00000272  kilowatt  hour 
0.0093          B.T.U. 


1  HORSE-POWER  HOUR  = 
0.746  kilowatt  hours 
1,980,000  foot-pounds 

2,545  B.T.U. 
273,740  kilogram  meters 

2.64  pounds  of  water  evaporated 
from  and  at  212°  F. 


1  HORSE  POWER  = 

746  watts 
0.746  kilowatts 
1.0136  cheval-vapeur 
33,000  foot-pounds  per  minute 
550  foot-pounds  per  second 
2,545  B.T.U.  per  hour 

42.4  B.T.U.  per  minute 
0.707  B.T.U.  per  second 
2.64  pounds    of    water    evaporated 
from  and  at   212°  F. 

1  FOOT-POUND  = 
1.356  joules 

0.1383  kilogram  meter 

0.000000377  kilowatt  hour 
0.001285        B.T.U. 
0.0000005      horse-power  hour 

1  B.T.U.  = 

1,055        watt  seconds 
778  foot-pounds 

107.6  kilogram  meters 

0.000293  kilowatt  hour 

0.000393  horse-power  hour 

0.001036  pound  water  evaporated  from, 
and  at  212°  F. 

1  CHEVAL-VAPEUR  = 
75  kilogrammeters  per  second 


0.9863  horse  power 
0.7357  kilowatt 


877 


APPENDIX  F. 

MISCELLANEOUS  CONVERSION  FACTORS. 


1   POUND  PER  SQUARE  INCH  = 

2.0355    inches  of  mercury  at  32°  F. 
2.0416    inches  of  mercury  at  62°  F. 
'2.309      feet  of  water  at  62°  F. 
0.07031  kilogram  per  square   centi- 
meter 

0.06804    atmosphere 
51.7          millimeters    of   mercury    at 
32°  F. 

1   FOOT  OF  WATER  AT  62°  F.  = 
0.433  pound  per  square  inch 
62.355  pounds  per  square  foot 

0.883  inch  of  mercury  at  62°  F. 
821.2      feet    of    air    at    62°  F.    and 
barometer  29.92 

1  INCH  OF  WATER  62°  F.  «= 

0.0361    pound  per  square  inch 
5.196      pounds  per  square  foot 
0.5776    ounce  per  square  inch 
0.0736    inch  of  mercury  at  62°  F. 
68.44        feet    of    air    at    62°  F.   and 
barometer  29.92 

1   FOOT  OF  AIR  AT  32°  F .  AND  BAROMETER 

29.92  = 

0.0761  pound  per  square  foot 
0.0146  inch  of  water  at  62°  F. 

1   INCH  OF  MERCURY  AT  62°  F.  = 
0.4912  pound  per  square  inch 
1.132    feet  of  water  at  62°  F. 
13.58      inches  of  water  at  62°  F. 


1   ATMOSPHERE  = 

760.0      millimeters   of   mercury   at 

32°  F. 

14.7      pounds  per  square  inch 
29.921  inches  of  mercury  at  32°  F. 
2,116.0      pounds  per  square  foot 

1.033  kilograms  per  square  centi- 
meter 

1  MILLIMETER  =  0.03937  inch 

1  CENTIMETER  =  0.3937  inch 

1  METER  =  39.37  inches 

1  METER  =  3.2808  feet 

1  SQUARE  METER  =  10.764  square  feet 

1   LITER  = 

61.023  cubic  inches 
0.264  U.  S.  gallons 

1   GRAM  = 

1  cubic  centimeter  of  distilled 

water 

15.43      grains  troy 
0.0353  ounce 

1   KILOGRAM  = 

2.20462  pounds  avoirdupois 


^878 


APPENDIX  G. 

RULES  FOR  FIREMEN  USING  ILLINOIS  AND  INDIANA  COAL  IN  HAND- 
FIRED  FURNACES. 

(Formulated  by  the  Coal  Stoking  and  Anti-Smoke  Committee  of  the  Illinois 
Coal  Operator's  Association.) 

1.  Break  all  lumps  and  do  not  throw  any  in  furnace  any  larger  than 
one's  fist.     The  reason  for  this  is,  that  large  lumps  do  not  ignite  promptly 
and  their  presence  also  causes  holes  to  form  in  the  fire,  which  allow  the 
passage  of  too  much  air. 

2.  Keep  the  ash  pits  bright  at  all  times.     If  they  become  dark  it  is 
evident  that  the  fire  is  getting  dirty  and  needs  cleaning,  which,  if  not 
done,  will  cause  imperfect  combustion  and  smoke.     If  the  furnace  is 
equipped  with  a  shaking  grate,  it  should  be  operated  often  enough  to 
prevent  any  accumulation  of  ashes  in  the  fire.     Do  not  allow  ashes  to 
collect  in  the  ash  pits,  as  they  not  only  shut  off  the  air  supply,  but  may 
cause  the  grate  to  be  burned. 

3.  In  firing  do  not  land  the  coal  all  in  one  heap,  but  spread  it  over 
as  wide  a  space  as  possible  as  it  leaves  the  shovel.     A  little  practice 
will  enable  one  to  catch  the  proper  motion  to  give  the  shovel  to  make 
the  coal  spread  properly. 

4.  Place  the  fresh  coal  from  the  bridge  wall  forward  to  the  dead  plate 
and  do  not  add  more  than  3  or  4  shovels  at  a  charge.    If  this  amount 
makes  smoke  it  should  be  reduced  till  smoke  ceases,  which  means,  of 
course,  that  firing  will  be  at  more  frequent  intervals  than  formerly  to 
keep  up  steam.     This  rule  applies  in  cases  where  the  boiler  is  worked 
at  a  large  capacity.     In  such  instances,  however,  where  a  small  capacity 
only  is  required,  firing  by  the  coking  method  is  the  best,  wherein  the 
fresh  coal  is  placed  at  the  front  of  the  fire  and  pushed  back  and  leveled 
when  it  has  become  coked. 

5.  Fire  one  side  of  the  furnace  at  a  time  so  that  the  other  side  contain- 
ing a  bright  fire  will  ignite  the  volatile  gases  from  the  fresh  charge. 

6.  Do  not  allow  the  fire  to  burn  down  dull  before  charging.     If  this 
is  done,  it  will  not  only  result  in  a  smoky  chimney,  but  an  irregular  steam 
pressure. 

7.  Do  not  allow  holes  to  form  in  the  fire.     Should  one  form,  fill  it  by 
leveling  and  not  by  a  scoop  full  of  coal.     Keep  the  fire  even  and  level 
at  all  times.     As  far  as  possible  level  the  fire  after  the  coal  has  become 

coked. 

879 


880  STEAM  POWER  PLANT  ENGINEERING 

8.  Carry  as  thick  a  fire  as  the  draft  will  allow,  but  in  deciding  on  th§ 
proper  thickness,  judgment  must  be  exercised.     If   the  draft  is  poor 
a  thin  fire  will  be  in  order,  but  if  strong,  a  thicker  fire  should  be  carried. 

9.  Regulate  the  draft  by  the  bottom  or  ash  pit  doors  and  not  by  the 
stack  dampers,  because  when  the  stack  damper  is  used  it  tends  to  pro- 
duce a  smoky  chimney,  as  it  reduces  the  draft,  while  the  closing  of  the 
ash  pit  door  diminishes  the  capacity  to  burn  coal.     If  strict  attention 
is  given  to  firing,  and  accounting  to  demand,  for  steam,  there  will  be  no 
occasion  to  have  recourse  to  dampers,  except  when  there  is  a  sudden 
interruption  in  the  amount  of  steam  being  used. 

10.  A  good  general  rule  is  to  fire  little  and  often,  according  to  steam 
demands,  rather  than  heavy  and  seldom.     The  former  means  economy 
in  fuel  and  a  clean  chimney,  while  the  latter  signifies  extravagance  in 
fuel  and  a  smoky  chimney. 


APPENDIX  H. 

MOLLIER'S   DIAGRAM. 

The  steam  tables  give  values  of  the  simultaneous  physical  properties 
of  steam,  such  as  pressure,  entropy,  temperature,  etc.  When  certain 
of  these  properties  are  known  the  remainder  can  be  obtained  from  the 
tables.  The  simultaneous  properties  can  also  be  shown  by  means  of  a 
diagram  each  point  on  which  represents  steam  in  a  perfectly  definite 
condition. 

Fig.  462  gives  a  skeleton  outline  of  such  a  diagram  and  Fig.  463  a 


B.T.j£p§r  Pound  of  Steam 
FIG.  462. 


reduced  reproduction  of  the  complete  chart  as  ordinarily  constructed. 
Referring  to  Fig.  462,  abscissas  represent  the  heat  contents  or  B.T.U. 
per  pound  of  steam  and  ordinates  represent  the  total  entropy.  Vertical 
lines  then  represent  lines  of  constant  heat  content,  and  horizontal  lines 
constant  entropy.  PlPl  and  P2P2  represent  lines  of  constant  pressure 
and  X1Xl  and  X2X2  lines  of  constant  quality.  Evidently  any  point  in 
the  chart  represents  a  fixed  condition  of  heat  content,  pressure,  quality 
and  entropy  as  determined  by  its  location  with  respect  to  the  different 
lines.  Thus  point  1  represents  a  pressure  Pl  as  determined  by  the 
numerical  value  of  line  P^P^  quality  xl  by  its  location  on  line  X^X^ 

881 


882  STEAM  POWER  PLANT  ENGINEERING 

entropy  nt  by  its  projection  Nl  on  the  F  axis,  and  heat  content  Hl  by 
its  projection  on  the  X  axis. 

The  principal  advantages  of  a  total  heat-entropy  diagram  over  the 
tables  are  that  they  give  the  properties  of  wet  and  superheated  steam 
and  offer  a  simple  means  of  solving  many  problems  without  calculations. 
For  example,  the  chart  offers  a  ready  solution  of  problems  involving 

(a)  Adiabatic  expansion. 

(b)  Throttling. 

(c)  Expansion  with  frictional  resistances. 

(a)  Adiabatic  Expansion:  From  thermodynamics  we  know  that  during 
an  adiabatic  change  the  entropy  is  constant;  thus,  in  expanding  from 
pressure  P1  and  condition  represented  in  point  1  to  a  lower  pressure  P2 
it  is  only  necessary  to  find  the  intersection  2  of  a  horizontal  line  from 
point  1  with  line  P2P2-  The  various  properties  corresponding  to  point 
2  can  be  read  directly  from  the  diagram. 

The    line    1-2  =  Ht  H2    represents    the    difference    in    heat    content 
following  adiabatic  expansion   from  pressure   Px  and  condition   1  to 
pressure  P,  or 
line  H1H2  =  Hl-  H2  =  z,^  +  ql  -  .r2r2  -  q,. 

The  quality  x2  is  read  directly  from  the  intersection  of  line  1-2  with 
the  constant  quality  line  X2X2. 

The  entropy  n2,  of  course,  remains  the  same. 

From  equation  (73),  p.  we  find  that  the  velocity  due  to  adiabatic 
expansion  is 

V  =  223.9  VH,-  H2. 

Mollier  has  added  along  the  margin  of  the  diagram  (Fig.  463)  a  scale 
of  velocity  so  that  V  may  be  ascertained  by  laying  off  the  length  H1 H2 
on  the  scale. 

Example:  Steam  at  120  pounds  absolute,  quality  0.98,  expands  adia- 
batically  to  a  back  pressure  of  2  pounds  absolute.  Find  the  quality 
and  heat  content  at  the  lower  pressure. 

From  Fig.  462  we  locate  Pt  at  the  intersection  of  pressure  curve  120 
and  quality  curve  0.98.  The  corresponding  values  of  Hl  and  nl  are 
found  by  interpolation  to  be  1174.7  and  1.564  repectively.  Follow 
horizontal  line  1.564  until  it  intersects  pressure  line  P2.  The  corre- 
sponding values  of  H2  and  x2  are  found  to  be  910  and  0.797  respectively. 
The  horizontal  intercept  between  the  two  pressure  lines  laid  off  on  the 
velocity  diagram  gives  V  =  3640  feet  per  second. 

Supposing  the  steam  to  be  superheated  200  degrees  instead  of  being 
wet,  find  the  quality  and  heat  content  at  the  end  of  expansion. 


E-'I 


E— I! 


~     3 


884  STEAM  POWER  PLANT  ENGINEERING 

Locate  Px  at  the  intersection  of  pressure  curve  120  and  superheat 
curve  200.  The  corresponding  values  of  Hl  and  nl  are  found  to  be 
1295  and  1.703  respectively.  Follow  horizontal  line  1.703  until  it  inter- 
sects pressure  line  P2.  The  corresponding  values  of  H2  and  x2  are 
found  to  be  990  and  877  respectively. 

(b)  Throttling:  If  steam  expands  through  a  small  orifice  without  the 
addition  or  abstraction  of  heat  and  is  brought  finally  to  its  initial  con- 
dition its  total  heat  will  be  unchanged.     This  process  is  called  throttling 
and  occurs  when  steam  passes  through  a  reducing  valve.     Vertical  lines 
in  Figs.  462  and  463  are  lines  of  constant  total  heat  and  consequently 
show  the  changes  in  the  condition  of  steam  which  result  from  throttling. 
Thus  in  throttling  steam  from  pressure  P1;  Fig.  462,  to  P2  it  is  only 
necessary  to  find  the  intersection,  4,  of  a  vertical  line  from  point  1  with 
line  P2P2. 

Example:  Steam  at  200  pounds  pressure  and  quality  0.96  passes 
through  a  reducing  valve  and  its  pressure  is  lowered  to  15  pounds. 
Find  its  quality  at  the  lower  pressure.  The  intersection  of  pressure  line 
190  with  quality  line  0.96  gives  #t  =  1165.  Follow  vertical  line  1165 
until  it  intersects  pressure  line  15.  The  corresponding  value  for  x2 
is  found  to  be  30,  that  is,  the  steam  is  superheated  30  degrees. 

To  what  pressure  must  the  steam  be  reduced  in  order  that  it  may  be 
dry  and  saturated?  Follow  vertical  line  1165  until  it  intersects  the 
saturation  curve.  The  corresponding  pressure  is  found  to  be  30  pounds. 

(c)  Expansion  Involving  Frictional  Resistances:  As  steam  expands  in 
the  nozzle  of  a  turbine  or  passes  between  the  vanes  it  experiences  fric- 
tional  resistances  which  cause  it  to  give  up  less  energy  than  it  would 
under  ideal  conditions.     The  work  of  friction  causes  the  entropy  of  the 
steam  at  its  lowest  temperature  to  be  greater  than  it  would  be  if  adia- 
batic  expansion  occurred  and  serves  to  increase  its  dryness  fraction. 

If  y  one  hundredths  of  the  heat  Ht  —  H2  (given  up  in  adiabatic 
expansion)  is  lost  due  to  friction,  the  heat  available  for  useful  work  is 

(!-?/)(#,-#,), 
the  resulting  velocity  of  the  jet  is 

V  =  223.9  V(l  -y)  (Hi  -Ht), 
and  the  increase  in  quality  of  the  exhaust  steam  is 

y  (#»-#,). 


These  equations  may  be  readily  solved  by  means  of  the  diagram. 
Referring  to  Fig.  462  line  1  3  represents  an  expansion  from  pressure 
t  to  pressure  P2  with  frictional  resistances. 


APPENDIX  H  885 

From  the  diagram 


12       HJIi 

(1  -  y)  (H,  -  #,)  =  line  1  N  =  H,HZ. 

Increase  in  quality  =  -  -  —  —  —  =  distance  2  3  between  X2  X2  and 

XSX3. 

Increase  in  entropy  =  N3  =  Nfl^ 

Example:  Steam  at  160  pounds  absolute  initial  pressure,  quality  0.97, 
expands  through  a  nozzle  to  a  back  pressure  of  2  pounds  absolute.  If 
15  per  cent  of  the  heat  energy  is  lost  in  friction,  find  the  quality  of  the 
steam  at  the  lower  pressure  and  the  velocity  of  the  jet. 

From  Fig.  462  we  locate  Pl  at  the  intersection  of  pressure  curve  160 
and  quality  curve  0.97.  The  corresponding  value  of  Hl  is  1170.  Follow 
line  1170  horizontally  until  it  intersects  pressure  line  P2.  From  the 
diagram  we  find  for  adiabatic  expansion  H2  =  910  x2  =  0.797.  But  the 
friction  increases  the  heat  content  at  the  end  of  expansion  an  amount 
0.15  X  Hi  -  H2  =  0.15  (1170  -  910)  =  39,  so  that  the  final  heat  con- 
tent =  910  +  39  =  949. 

Follow  pressure  line  P2  until  it  intersects  heat  line  949.  The  quality 
x2  is  found  to  be  0.836  and  the  entropy  n2  =  1.632.  From  the  velocity 
scale  we  find  V  =  3320  feet  per  second  for  /^  -  H2  =  (1170  -  949). 


INDEX 


Absorption  dynamometer,  857. 

Acetylene,  properties  of,  26. 

Acidity,  tests  for,  in  oils,  674. 

Acme  bucket  trap,  590. 

Acton  atmospheric  relief  valve,  666. 

Ados  CO2  recorder,  743. 

Aero-pulverizer  powdered  coal  burner,  50. 

Air  chambers,  528. 

Air-cooled  surface  condensers,  428. 

Air  lift,  572. 

Air,  properties  of,  26. 

Air  pumps,  552-560. 

size  of  dry,  558. 

size  of  wet,  553. 

Air  required  for  operating  air  lift,  573. 
Air  required  for  combustion,  28. 
Air  spaces,  grate  bars,  114. 
Air  supply  above  grate,  151. 
Air  thermometers,  recording, 
Air  vs.  steam  as  an  oil  atomizer,  60. 
Alarm,  high  and  low  water,  120. 
Alberger  barometric  condenser,  412. 

cooling  tower,  459. 

rotative  dry  air  pumps,  557. 
Alden  absorption  dynamometer,  857. 
Allis-Chalmers  steam  turbine,  372. 
Alternate  method  of  starting  and  stopping 

boiler  tests,  826. 

American  underfeed  stoker,  139. 
Analyses  of  boiler  scales,  473. 

of  flue  gases, 

of  fuel  oils,  52. 

of  typical  American  coals,  23. 

of  waters  for  boiler  feeding,  473. 
Anchors,  pipe,  622. 

Anderson    automatic    non-return    valve, 
658. 

feed-water  pumping  system,  482. 

triple-duty  emergency  valve,  659. 
Animal  fats  and  oils,  669. 
Anthracite  coals,  15. 


Aqueous  vapor,  pressure  of,  399. 
effect  of,  on  degree  of  vacuum,  405. 

Armour  Glue  Works,  vacuum  ash  system 
at,  198. 

Armour  Institute,  brick  chimney  at,  227. 

Arndt's  econometer,  742. 

Ash  bins,  182. 

Ash  conveyor,  vacuum  system,  196. 

Ash,  influence  of,  on  fuel  value  of  dry 
coal,  40. 

Ash-handling  systems,  181-204. 

Ash,  treatment  of,  in  boiler  tests, 

A.S.M.E.  rules  for  conducting  boiler  trials, 
822-845. 

A.S.M.E.    rules    for    conducting    engine 
tests,  846-872. 

Atmospheric  heaters,  486. 

Atmospheric  surface  lubrication,  675. 

Atmospheric  relief  valves,  666. 

Augmenter,  Parsons  vacuum,  444. 

Aurora  and  .Elgin  Interurban  Ry.,  coal- 
handling  system,  194. 

Austin  steam  separator,  579. 

Automatic  cut-off  vs.  throttling  engines, 
310. 

Automatic  injectors,  547. 

Automatic  non-return  valves,  658. 

Automatic  temperature  control,  646. 

Auxiliaries,   power  consumption  of  con- 
denser, 449. 

Auxiliaries,  measurement  of  steam  used 
by,  851.' 

Babcock  &  Wilcox  boilers,  81. 

chain  grate,  128. 

superheater,  163. 
Back  connection,  return   tubular  boiler, 

79. 

Back  pressure  on  engines,  283. 
Back-pressure  valves,  665. 
Baffle-plate  steam  separator,  579. 


887 


888 


INDEX 


Bagasse  as  fuel,  20. 
Balanced-draft  system,  264. 
Baragwanath  feed-water  heater,  495. 
siphon  condenser,  408. 
surface  condenser,  416. 
Barnard- Wheeler  cooling  tower,  457. 
Basement    plan,    West    Albany    station, 

N.  Y.  C.  R.  R.,  720. 
Bearings,  lubrication  of,  675. 
Belliss  engines,  tests  of,  with  superheated 

steam,  314. 
Belt  conveyors,  192. 
Bends,  pipe,  619. 
Bibliography: 

Cost  of  electric  power,  724. 

Cost  of  gas  power,  726. 

Cost  of  steam  power,  727. 

Cost  of  water  power,  728. 

Description  of  gas-driven  power  plants, 

798. 
Description  of  central  stations,  steam 

engines,  802-818. 
Description  of  central  stations,  steam 

turbines,  808,  819. 
Description  of  hydraulic  power  plants, 

798,  818. 

Description  of  isolated  station,  809. 
apartment  buildings,  809. 
manufacturing  plants,  810. 
office  buildings,  812. 
stores,  814. 

Design  of  power  plants, 
Binary- vapor  engines,  321. 
Bituminous  coals,  16. 
Blades,   arrangement  of,   in  steam  tur- 
bines, 352,  367. 
Blake  jet  condenser,  403. 
Blast  furnace  gas,  properties  of,  67. 
Bloomsburg  steam  jet,  246. 
Blowers,  fan,  249. 
tests  of,  257. 
steam  jet,  245. 
Blow-off    piping,    South    Side    Elevated 

R.R,  662. 
Blow-offs,  116. 
Blow-off  tank,  117. 
Blow-off  valve,  661. 
Boiler  compounds,  476. 
Boiler-feed  pumps  (see  Pumps). 
Boiler  room  area,  86. 
Boiler  tests,  A.S.M.E.  code,  822-845. 


Boiler    tests,  discrepancy  between  com- 
mercial and  experimental  results,  842. 
Boilers,  66-123. 

Babcock  &  Wilcox,  chain-grate,  128. 

Babcock  &  Wilcox,  hand-fired,  80. 

capacity  of,  104. 

classification  of,  68. 

cost  of,  112. 

efficiency  of,  98,  762. 

fire-box,  71. 

furnaces  for,  68-88,  124-151. 

grates  for,  114,  126. 

heating  surface  of,  92. 

Heine,  83. 

horizontal  return  tubular,  74. 

horse  power  of,  93. 

inspector's    report    (1907),    Hartford 
Boiler  Insurance  Company,  474. 

Manning  vertical,  70. 

Parker  boilers,  85. 

performances  of,  99. 

Robb-Mumford,  73. 

Scotch-marine,  72. 

selection  of  type,  112.- 

settings  for,  68-90,  126-140. 

specifications  for,  754. 

Stirling,  87. 

vertical  tubular,  69. 

Wickes,  84. 

Booth  fuel  oil  burner,  56. 
Boston  Elevated,  cost  of  operation,  710. 
Brake  horse  power,  855. 
Brake,  rope,  856. 
Branch  fuel  oil  burner,  57. 
Brass  pipes,  608. 
Breeching,  240. 
Brick  chimneys,  224. 
Bristol  recording  air  thermometers,  737. 

thermo-electric  pyrometer,  737. 
Bucket  conveyor,  184. 
Bucket  traps,  590. 
Buckeye  skimmer,  118. 
Bundy  steam  separator,  579. 
Bunkers,  coal,  182. 
Burgeon,    specific    heat    of    superheated 

steam,  157. 

Burke 's  smokeless  furnace,  151. 
Burners,  oil,  53-61. 

powdered  coal,  44-51. 
Burnham  steam  meter,  734. 
Burning  point,  oils,  674. 


INDEX 


889 


Bui-sting  strength  of  pipes,  607. 
By-pass  system  of  piping,  625. 

Calorific  value  of  coals,  31. 
Calorimeters,  fuel,  747. 
Calorimeters,  steam,  745. 
Cannel-coal  gas,  properties  of,  67. 
Capacity,  effect  of,  on  boiler  efficiency,  104. 
Carbon  dioxide,  properties  of,  26. 

percentage  of,  in  flue  gases,  30. 
Carbon    monoxide,    heat    losses   due   to 
formation  of,  36. 

properties  of,  26. 

Carbu retted  water  gas,  properties  of,  67. 
Carnot  cycle,  267. 

Carpenter  separating  calorimeter,  745. 
Cast-iron  pipes,  607. 
Central  condensing  systems,  439. 
Central   hydrostatic   cylinder  lubricator, 

684. 

Centrifugal  oilers,  678. 
Centrifugal  pumps,  560-567. 

characteristics  of,  565. 

performance  of,  563. 

tables  of  sizes,  566,  567. 

types  of,  560. 

Centrifugal  steam  separators,  578. 
Chain  grates,  126. 
Chattanooga  Electric   Company,  test  of 

spray  fountain,  455. 
Check  valves,  660. 

Chemical  purification  of  feed  water,  476. 
Chicago  setting,  hand-fired  furnace,  143. 
Chimney  at  Armour  Institute,  227. 
Chimney  draft,  207. 
Chimney    draft,    table    of,    for    various 

temperatures,  210. 
Chimney,  efficiency  of,  241. 

height  of,  for  burning  fuel  oil,  218. 

test  of  100-foot  steel,  213. 
Chimney  vs.  mechanical  draft,  2.61. 
Chimneys,  207-244. 

brick,  224. 

classification  of,  218. 

core  and  lining  for,  230. 

cost  of,  243. 

Custodis  radial  brick,  225,  234. 

dimensions  of,  216,  244. 

formulas  for,  212. 

foundations  for,  240. 

guyed  steel,  219. 


Chimneys,  materials  for  brick,  230. 

self-sustaining  steel,  219. 

stability  of  brick,  231. 

stability  of  steel,  223. 

steel,  219. 

strain   sheet   for   reenforced    concrete, 
234. 

thickness  of  walls,  brick,  226. 

thickness  of  shell,  steel,  220. 

Weber  reenforced  concrete,  235. 
Cincinnati  Traction  Company,  coal  con- 

veyors,-195. 
Circulating  pumps,  572. 
Classification  of  boilers,  68. 

chimneys,  218. 

condensers,  400. 

feed-water  heaters,  485. 

fuel-oil  burners,  53. 

fuels,  14. 

lubricating  oils,  669. 

powdered-coal  burners,  44. 

pumps,  522. 

steam  separators,  576. 

steam  traps,  588. 

steam  turbines,  327. 

stokers,  126. 

testing  instruments,  730-749. 
Clearance  volume,  influence  of,  on  engine 

economy,  281. 
Coal,  15. 

analysis  of,  for  boiler  tests,  829. 

anthracite,  15. 

bituminous,  16. 

calorific  value  of,  31. 

composition  of,  23. 

measurements  of,  boiler  tests,  852. 

powdered,  44-51. 

proximate  analysis,  31. 

purchasing,  42. 

sampling,  828. 

specifications  for  purchasing,  769. 

storage  of,  181. 

ultimate  analysis,  31. 

washed,  40. 

Coal  and  ash  handling,  181-204. 
Coal  bunkers,  182. 
Coal  gas,  properties  of,  67. 
Coal  fields  of  the  United  States  (ref.\  16. 
Coal  hoppers,  202. 
Coal  valves,  205. 
Cochrane  heater,  487. 


890 


INDEX 


Coefficient  of  expansion,  pipe  materials, 

620. 

Coke-oven  gas,  properties  of,  67. 
Cold  test,  oils,  674. 
Columbia  expansion  trap,  592. 
Combustion,  24. 
Commercial     National     Bank     Building, 

Chicago,  ash  system,  192. 
Commonwealth    Edison    Company,    Fisk 

Street  Station,  774-787. 
Compressed-air  oiling  system,  681. 
Compressed  air,  power  required,  air  lift, 

572. 
Compression,  effect  of,  on  engine  economy, 

283. 

Compound  engines,  300. 
Compounds,  boiler,  476. 
Condensers,  397-469. 

air  for  cooling  purposes,  surface,  429. 

Alberger  barometric,  412. 

Baragwanath  siphon,  408. 

barometric,  411. 

Blake  jet,  403. 

choice  of,  450. 

classification  of,  400. 

cooling  water  for,  404. 

cost  of,  450. 

counter-current,  411. 

dry  tube,  424. 

economical  vacuum  for,  451. 

ejector,  410. 

extent  of  cooling  surface  for,  421. 

function  of,  398. 

high-vacuum,  441. 

independent,  434. 

injection  orifice,  407. 

jet,  401. 

Korting  multi-jet,  446. 

location  of,  433. 

multi-flow  surface,  419. 

Schutte  ejector,  410. 

siphon,  408. 

sizes  of  siphon,  409. 

specifications  for, 

surface,  dry  air-cooled,  428. 

surface,  evaporative,  433. 

surface,  water-cooled,  416. 

tests  of  surface,  434. 

Tomlinson  barometric,  415. 

volume  of  condenser  chamber  for  jet, 
408. 


Condensers,  Weighton  multi-flow,  419. 

Weiss  barometric,  411. 

Westinghouse-Leblanc,  445. 

Wheeler  admiralty,  417. 

Worthington  barometric,  414. 
Condensers,  Worthington  jet,  402. 
Concrete  chimneys,  234-240. 
Condensation  and  leakage  losses  in  en- 
gines, 279. 

Condensing,    influence    on    engine    econ- 
omy, 307. 

Condensing  plant,  elementary,  7. 
Condensing  plant  with  full  complement 

of  heat-saving  appliances,  10. 
Conoidal  pump,  test  of,  568. 
Conversion  tables,  878. 
Conveyors,  183-197. 
Cooling  ponds,  454. 
Cooling  towers,  456. 

Cooling  towers,  fan  vs.  natural  draft,  460. 
Copper  pipes,  607. 
Corliss  engine,  269. 
Correction   factors   for   steam    turbines, 

387. 
Cost  of  boilers  and  settings,  112. 

chimneys,  243. 

condensers,  450. 

engines,  326. 

evaporating  water,  105. 

handling  coal  and  ashes,  201. 

mechanical  draft  systems,  263. 

pipe  flanges,  614. 

power  (see  power  costs). 

stokers,  151. 

turbines,  392. 
Costs,  operating,  693. 
Coverings  for  steam  pipes,  616. 
Crusher  and  cross  conveyor,  190. 
Curtis  steam  turbine,  350. 
Curve  load  factor,  691. 
Custodis  radial  brick  chimney,  234. 
Cut-off,  commercial,  860. 
Cut-off,  point  of,  861. 
Cylinder  condensation,  279. 
Cylinder  cups,  682. 
Cylinder  lubrication,  682. 
Cylinder  ratios,  compound  engines,  300. 

Damper  regulators,  118. 
Davis  back-pressure  valve,  665. 
Dean  air  pump,  552. 


INDEX 


891 


De  Laval  centrifugal  pump,  568. 

steam  turbine,  331. 
Density  of  air  and  flue  gas,  209. 
Depreciation  of  powdered-coal  f  urnace,44. 
Depreciation,  rate  of, 
Depreciation  percentages,  Chicago  Trac- 
tion Valuation  Commission,  696. 
Desmond  injector,  test  of,  549. 
Detroit  Edison  Co.,  coal-handling  system, 

196. 

Diagram  factor,  steam  engine,  862. 
Diaphragm  valve,  647. 
Differential  traps,  594. 
"Direct"  steam  separator,  580. 
Disk  water  meter,  732. 
Divergent  nozzle,  design  of,  334. 
Dodge,   A.    R.,   specific  heat    of    super- 
heated steam,  156. 

Double-deck  turbine  installation,  381. 
Double-flow  steam  turbine,  369. 
Double  stoker,  135. 
Down-draft  furnace,  139. 
Draft,  balanced,  264. 

chimney,  207. 

for  powdered-coal  burner,  47. 

forced,  251. 

gauges,  735. 

influence  of,  on  boiler  efficiency,  106. 

induced,  252. 

mechanical,  245-266. 
Drainage  of  jackets  and  receivers,  597. 
Drains,  office  building,  603. 
Drips,  586. 

high-pressure,  588. 

low-pressure,  586. 

removal  of  oil  from,  587. 

under  alternate  pressure  and  vacuum, 
599. 

under  vacuum,  598. 
Dry-air  pumps,  537. 
Dry-air  surface  condensers,  428. 
Dry  docks,   centrifugal  pump  character- 
istics for,  565. 

Dry  tube  surface  condensers,  424. 
Dulong's  formula,  32. 
Dunham  steam  trap,  593. 
Duplex  coal  valve,  204. 
Duplex  steam  pump,  524. 
Duplicate  piping  system,  624. 
Dutch  oven,  141. 
Duty,  pump,  536. 


Economizers,  508. 

factors  for  determining  installation  of, 
532. 

Green,  510. 

heat  transmission  in,  512. 

tests  of,  515. 
Edwards  air  pump,  555. 
Efficiencies  of  boilers  and  grates,  98. 

boilers  with  oil  fuel,  54. 
Efficiencies  of  boiler-feed  pumps,  534. 

centrifugal  pumps,  565-568. 

compound    engines,    saturated    steam,. 
306. 

compound  engines,  superheated  steam, 
317,  321. 

fans,  257. 

piston  pumps,  533. 

simple  engines,  saturated  steam,  296. 

simple  engines,  superheated  steam,  316. 

steam  turbines,  384. 

triple-expansion    engines,    superheated 
steam,  318. 

triplex  pumps,  544,  545. 
Efficiency,  air  lift,  574. 

Carnot  cycle,  267. 

condensing  plants,  9,  13. 

furnace,  99. 

grate,  99. 

mechanical,  275. 

non-condensing  plants,  4. 

Rankine  cycle,  268. 

thermal,  273. 

Ejector  condenser,  410,  446. 
Ejector,  Shone,  603. 
Electrical  power,  cost  of,  704-729. 
Elementary  condensing  plants,  2. 
Elementary  non-condensing  plant,  7. 
Elementary  theory,  Curtis  turbines,  358. 

De  Laval  turbine,  333. 

Westinghouse-Parsons  turbine,  373. 
Elevating  tower,   cable-car  distribution^ 

193. 
Elevating   tower,   hand-car  distribution, 

196. 

Ellison's  universal  steam  calorimeter,  746. 
Emergency  valves,  658. 
Engines  (steam),  267-326. 

A.S.M.E.  code  for  testing,  846-872. 

automatic  cut-off,  310. 

back  pressure,  effect  of,  on  economy, 
283. 


892 


INDEX 


Engines  (steam),  binary  vapor,  321. 

clearance  volume,  281. 

compound,  300. 

compression,    effect    of,    on    economy, 
283. 

condensing,  effect  of,  on  economy,  307. 

cost  of,  326. 

cylinder  condensation,  279. 

economy  of  (see  Tests). 

efficiencies  of  (see  Efficiencies). 

friction  of,  284. 

heat  losses  in,  278. 

high-speed,  290. 

ideal,  267. 

incomplete  expansion,  loss  due  to,  282. 

increasing    initial    pressure,   effect    of, 
286. 

jackets,  influence  of,  170. 

leakage  losses  in,  279. 

low-speed,  299. 

mechanical  efficiency,  275. 

non-condensing,  test  of,  296. 

receiver-reheaters,  economy  of,  287. 

simple,  291. 

single-acting,  290. 

specifications,  750. 

sulzer,  319. 

superheated  steam,  313. 

tests  of  (see  Tests). 

thermal  efficiency  of,  273. 

throttling  vs.  automatic  cut-off,  310. 

triple  and  quadruple  expansion,  305. 

wire  drawing,  effect  of,  284. 

with  low-pressure  turbines,  376. 
Entropy  diagram,  863,  881. 
Equation  of  pipes,  640. 
Exhaust  heads,  585. 
Exhaust  piping,  642. 
Expansion  traps,  592. 
Expansion  of  pipe  materials,  618. 
Expansion,  ratio  of,  862. 
Extended  boiler  front  setting, 
Evaporation,  cooling  pond,  454. 

cost  of,  coal  fuel,  105. 

cost  of,  oil  vs.  coal,  55. 

from  and  at  212°  F.  per  square  foot  per 
hour,  95. 

rate  of,  boilers,  96. 

rates  of,  in  still  air,  454. 

unit  of,  88. 
Evaporative  surface  condenser,  433. 


Factor  of  evaporation,  88, 

Fan  draft,  249. 

Fans,  capacity  of  induced-draft,  262. 

capacity  of  forced-draft,  261. 

performance  of,  258. 

theory  of,  252. 
Feed  water,  analyses  of,  473. 
Feed- water  heaters  (see  Heaters). 
Feed-water  heating  system,  choice  of,  516. 
Feed-water  piping,  647. 
Feed- water  purification,  471. 
Feed-water  regulators,  542. 
Fery  radiation  pyrometer,  740. 
Filters,  oil,  688. 
Fire-box,  boilers,  71. 
Fire-tile  "Economy,"  131. 
Fire-tile  combustion  chamber,  128. 
Fire,  thickness  of,  110. 
Fire-tube  boiler,  71. 
First  National  Bank  Building,  Chicago, 

power  costs,  710. 
Fisher  pump  governor,  541. 
Fittings,  pipe,  610. 
Fixed  carbon  in  coat  18. 
Fixed  charges,  693. 
Flanged  fittings,  610. 
Flanges,  table  of  extra  heavy,  615. 
Flanges,  table  of  standard,  614. 
Flap  coal  valve,  205. 
Flash  point,  oil  testing,  673. 
Flemming  four- valve  engine,  320. 
Flinn  trap,  595. 
Float  trap,  589. 
Floor    space,    turbine    vs.    reciprocating 

engine,  383. 
Flow  of  steam  in  pipes,  632. 

steam  through  nozzles,  336. 

water  through  pipes,  650. 
Flue-gas  analysis,  29,  741. 
Flue-gas  apparatus,  Ados  recording,  743. 

Arndt's  econometer,  indicating,  742. 

Orsat,  741. 

Sarco  recording,  744. 
Flush-front  boiler  setting, 
Fly-wheel  pumps,  523. 
Foot  valves,  668. 
Forced  draft,  245-266. 
Forced-feed  lubricator,  684. 
Forcing  capacity  of  boilers,  109. 
Foster  back-pressure  valve,  665. 

pressure  regulator,  667. 


INDEX 


893 


Foster  superheater,  165. 

Foundations,  chimney,  240. 

Fountain,  spray,  455. 

Four- valve  engines,  tests  of,  298,  306. 

Friction  of  engines,  284. 

Friction  of  water  in  pipes,  652. 

Friction  tests  of  oil,  674. 

Friction  through  valves  and  fittings,  639, 
653. 

Fuel,  cost  of,  700. 

Fuel  calorimeters,  747. 

Fuel  oil,  51-66. 

Fuel-oil  burners  (see  Burners). 

Fuels  and  combustion,  14-16. 

Fuels,  classification  of,  14. 

Function  of  the  condenser,  398. 

Furnace  arch  bars,  79. 

Furnace  efficiency,  99. 

Furnace  temperature,  influence  on  boiler 
efficiency,  111. 

Furnace  influence  on  gas  composition,  36. 

Furnace  for  burning  oil  fuel,  59. 

Furnace  for  burning  powdered  coal,  47. 

Furnace,  smokeless  (see  Smokeless  fur- 
naces). 

Fusible  plugs,  121. 

Gaseous  fuels,  66. 

characteristics  of,  67. 
Gauge  cocks,  3,  119. 
Gauges,  water,  119. 
Gate  valves,  655. 
Geipel  steam  trap,  593. 
Globe  valves,  655. 
Government  specifications  for  purchasing 

coal,  769. 

Governor,  steam  pump,  541. 
Goubert  feed-water  heater,  492. 
Grate,  loss  of  fuel  through,  37. 
Grate  surface,  95,  98. 
Grate  bars,  thickness  of,  114. 
Grates,  chain,  126. 

rocking,  116. 

stationary,  114. 
Gravity  oil  feed,  680. 
Gravity,  Baume  oils,  672. 
Gravity,  specific  oils,  672. 
Grease  extractor,  584. 
Greases,  671. 
Green  chain  grate,  126. 

economizer,  510. 
Guyed  steel  chimneys,  219. 


Hamilton-Holzworth  turbine,  362. 

Hamler-Eddy  Smoke  Recorder, 

Hammel  fuel  oil  burner,  57. 

Hancock  injector,  546. 

Hand  shoveling,  183. 

Hangers,  for  pipes,  622. 

Hartford     Boiler     Insurance     Company, 

annual  report  (1907),  474. 
Hartford  boiler  specifications,  754. 
Hawley  down-draft  furnace,  139. 
Headers,  main  steam,  563. 
Heat  balance,  boiler  tests,  832. 
Heat  distribution,  condensing  plants,  9, 

13. 
Heat  distribution,  non-condensing  plants, 

4. 

Heat  losses  in  burning  coal,  32. 
in  the  chimney  gases,  33. 
in  steam  engines,  278. 
Heat  transmission,  boilers,  90, 
closed  heaters,  497. 
economizers,  512. 
influence  of  scale  on,  472. 
superheaters,  170. 
Heating  surface,  boilers,  92,  95. 
Heaters,  feed- water,  471-521. 
Baragwanath,  595. 
choice  of,  516. 
classification  of,  485. 
closed,  486. 
Cochrane,  487. 
counter-current,  485. 
flue-gas,  485. 
Goubert,  492. 
Harrisburg,  494. 
Hoppe's,  489. 
induced,  486,  506. 
live  steam,  485-507. 
open,  486,  504. 
Otis,  494. 

parallel-current,  492. 
primary,  486. 
secondary,  486. 
single-flow,  492. 
steam  tube,  494. 
through,  486,  505. 
vacuum,  485,  506. 
Wainwright,  493. 
Webster,  488. 

Heater  and  purifier  combined,  489. 
Heine  boiler,  83. 


S94 


INDEX 


Heinrich  smokeless  furnace,  249. 

Heintz  expansion  trap,  594. 

Herringbone  grate,  115. 

Hewes  and  Phillips  air  pump,  556. 

Heyworth    Building,    Chicago,    plan    of 
piping,  421. 

High  and  low  speed  engines,  290. 

High-pressure  drips,  588. 

High-speed  double- valve  engine,  297. 

High-speed  single- valve  engine,  290. 

Hollow  bridge  wal,  151,  247. 

Holly  loop,  602. 

Hoppe's  feed-water  heater,  489. 
steam  separators,  577. 

Horizontal  return  tubular  boilers,  74. 

Horse  power  of  boilers,  93. 

Hot-well     temperatures,     surface     con- 
densers, 426. 

Hot-well  pumps,  560. 

Hunt  coal  conveyor,  189. 

Hydraulic  packing,  530. 

Hydraulic   Oil   Storage  Company's  fuel 
oil  system,  64. 

Hydraulic  valve  gear,  Curtis  turbine,  355. 

Hydrogen,  properties  of,  26. 
losses  due  to,  38. 

Hydrometer,  Baume",  oils,  672. 

Hydrostatic  cylinder  lubricator,  683 

JHygrometry,  468. 

Ideal  engine,  267. 

Illinois    Engineering    Company's    auto- 
matic vacuum  valve,  643. 

Impulse  turbine,  331. 

Incomplete  combustion,  loss  due  to,  36. 

Incomplete  expansion,  loss  due  to,  282. 

Increasing  boiler  pressure,  economy  of, 
286. 

Increasing  rotative  speed,  economy  of, 
290. 

Increasing  degree  of  vacuum,  cost  of,  451 . 

Increasing  degree   of   vacuum,   economy 
of,  308,  395. 

Identification  of  oils,  672. 

Independent  condensers,  434. 

Independently  fired  superheaters,  166. 

Indicated  horse  power,  853. 

Indicator  cards,  air  pump,  560. 

Indicator  cards,  analysis  of,  859. 
automatic  cut-off  engines,  311. 
four- valve  engine,  861. 


Indicator  cards,  throttling  engine,  311. 

Westinghouse-Parsons  turbine,  365. 
Indicator  springs,  tests  of,  854, 
Induced  draft,  252. 
Induced  heaters,  506. 
Initial  condensation,  279. 
Injection  orifice,  408. 
Injectors,  545-548. 

performance  of,  548. 

range  in  working  pressures,  350. 

vs.  steam  as  boiler  feeders,  550. 
Intermittent  oiling,  675. 
International    Gas    Company's    fuel    oil 

system,  63. 
Interest  charges,  693. 
Isolated  stations,  cost  of  power  in, 71 6-723. 
Isolated  stations,  influence  of  load  factof 
on  economy,  720. 

Jackets,  influence  of,  170. 

Jackets,  methods  of  draining,  597. 

Jet  condensers, 

Jet,  Bloomsburg,  246. 

Jet,  ring  steam,  246. 

Jets,  steam  consumption  of,  248. 

Jones  underfeed  stoker,  72. 

Kent's  wing-wall  furnace,  149. 

Kerosene,  use  of,  in  boilers,  477. 

Kerr  steam  turbine,  346. 

Keystone  separator,  578. 

Kieley  reducing  valve,  667. 

Kindling  temperatures,  25. 

Kirk  wood  oil  burner,  58. 

Kitts  feed-water  regulator,  542. 

Kitts  hydraulic  damper  regulator,  53. 

Knoblauch    and    Linde,   specific  heat  of 

superheated  steam,  156. 
Knowles  triplex  pump,  test  of,  544. 
Korting  fuel-oil  burner,  56. 
Korting  multi-jet  condenser,  446. 

Labor,  cost  of,  in  power  plants,  699. 

cost  of,  in  street  railway  plants,  701. 

cost  of,  in  tall  office  buildings,  702. 
Lea-Degan  three-stage  turbine  pump,  562, 

569. 

Leakage  of  steam  in  engines,  279. 
Leyland  automatic  cylinder  cup,  683. 
Life  of  power  plant  appliances,  694. 
Lignite,  18. 
Limit  of  superheat,  154. 


INDEX 


895 


Link    Belt  Company  coal-handling  sys- 
tem, 184. 

Live-steam  feed-water  heaters,  507. 
Load  factor,  691. 

influence  of,  on  cost  of  power,  692. 
Location  of  condensers,  433. 

of  separators,  580. 

of  traps,  596. 

Loew  grease  extractor,  584. 
Loop  header,  624. 
Loop,  Holly,  602. 
Loop,  steam,  600. 
Loss  of  heat  from  bare  pipes,  615. 
Loss  of  heat  from  covered  pipes,  616. 
Losses  in  burning  fuel,  32-40. 
Losses  in  steam  engines,  278. 
Low-pressure  drips,  586. 
Low-pressure  turbines,  376. 
Low-speed  engines,  299. 
Lubricants,  669-680. 
Lubrication,  atmospheric,  675. 

central  hydrostatic  system,  684. 

compressed-air  feed,  680. 

cylinder  lubrication,  682. 

forced  system,  684. 

gravity  systems,  680. 

Siegrist  system,  685. 
Lubricating  oils,  classification  of,  670. 

properties  of,  676. 

specific  gravity  of,  672. 
Lubricators,  hydrostatic,  683. 
Ludlow  angle  valve,  667. 
Lunkenheimer  sight-feed  lubricator,  684. 

Mahler  bomb  calorimeter,  747. 

Mains,  steam,  622. 

Maintenance,  699. 

Manning  vertical  boilers,  70. 

Marks'  and  Davis'  steam  tables, 

Marsh  gas,  properties  of,  26. 

Marsh  steam  pump,  527. 

Marsh  steam  pump,  test  of,  534. 

Materials  for  brick  chimneys,  230. 

Materials  for  pipes  and  fittings,  606. 

Materials  for  superheaters,  170. 

McClaves  argand  blower,  246. 

McDaniel  float  trap,  589. 

Mean    temperature     difference,    heaters, 

497. 
Measurement  of  heat  units  consumed  by 

engines,  849. 


Measurement  of  feed-water  consumption, 

engine,  849. 
Measurement  of  steam  used  by  auxiliaries, 

850. 

Mechanical  draft,  245-266. 
Mechanical  efficiency  of  engines,  275. 

of  pumps,  533. 

Mechanical  boiler-tube  cleaner,  121. 
Mechanical  purification  of  feed  water,  481. 
Mechanical  stokers,  125. 
Mechanical  valve    gear,  Curtis  turbine, 

354. 

Medium  and  low  speed  engines,  299. 
Mesh  steam  separators,  580. 
Meters,  steam,  734. 
Meters,  water,  732. 
Meyers  Bagasse  furnace,  21. 
Meyers  tan  bark  furnace,  22. 
Mineral  oils,  670. 
Mixed  pressure  turbines,  376. 
Moisture  in  air,  loss  due  to,  in  combustion, 
37. 

in  fuel,  loss  due  to,  37. 

in  steam,  determination  of,  827. 

in  steam,  effect  on  engine  economy, 
278. 

in  steam  evaporated  by  throttling,  312. 
Mollier  diagram, 
Mullan  valveless  air  pump,  556. 
Murgue's  theory,  centrifugal  fans,  254. 
Murphy  furnace,  137. 

Napier's  rule  for  the  flow  of  steam,  336. 
Natural-draft  cooling  tower,  460. 
Natural  gas,  properties  of,  67. 
Naval    Liquid    Fuel    Board,    report    of 

United  States,  65. 
Nitrogen,  properties  of,  26. 
Non-condensing  engines,  test  of,  296,  315. 
Non-condensing  plants,  arrangement  of, 

2,  5. 

exhaust  piping  in,  Paul  system,  643. 
exhaust  piping  in,  Webster  system,542. 
feed-water  piping  in,  647. 
open  heater  in,  506. 
Non-return  valves,  658. 
Norfolk  Traction  Co.'s  ash -handling  sys- 
tem, 200. 
Northwestern    Elevated    R.    R.    power 

house,  condenser  piping,  446. 
Nozzles,  De  Laval  steam  turbine,  332. 


896 


INDEX 


Nozzles,  flow  of  steam  through,  334. 

flow  of  water  through,  650. 

Kerr  steam  turbine,  346. 

theoretical  design  of  divergent,  334. 
Nugent  telescopic  oiler,  677. 

Office  buildings,  cost  of  power  in,  703- 

720. 

Oil  bath,  675. 
Oil  burners  (see  Burners). 

cups,  677. 

eliminators,  581. 

filters,  687. 

fuels,  analysis  of,  31. 

pressure  in  fuel  oil  systems,  63. 

separators,  581. 

storage,  64. 
Oil,  waste,  and  supplies,  cost  of,  in  power 

plants,  703. 
Oiler,  centrifugal,  678. 

gravity,  679. 

pendulum,  678. 

ring,  678. 

telescope,  677. 

Oiling  systems  (see  Lubricating  systems). 
Oils,  animal,  669. 

distilled,  670. 

identification  of,  672. 

mineral,  670. 

properties  of,  671. 

specific  gravity  of,  672. 

specifications  for,  672. 

tests  for,  672. 

Olefiant  gas,  properties  of,  26. 
Open  heaters,  486. 
Open  heater  vs.  closed  heater,  504. 
Operating  costs, 
Operating  costs,  reciprocating  engine  vs. 

steam  turbine,  392. 
Optical  pyrometers,  738. 
Orsat  apparatus,  741. 
Orifice,  size  of  injection,  408. 
Orifices,  flow  of  steam  through,  336. 

flow  of  water  through,  584. 
Otis  feed-water  heater,  428. 
Overhead  storage,  bucket  hoist,  195. 
Overload  capacity,  steam  turbines,  364. 
Oxygen,  properties  of,  26. 

Pan  surface,  open  heaters,  491. 
Parallel -current  condenser,  401. 
Parallel-current  feed-water  heater,  492. 


Parker  boiler,  85. 

Parr  coal  calorimeter,  748. 

Parsons  smokeless  furnace,  248. 

Parsons  vacuum  augmenter,  444. 

Paul  exhauster,  645. 

Paul  heating  system,  643. 

Peat,  18. 

Penberthy  injector,  546. 

Pendulum  oiler,  678. 

Pennel   saturated-air  surface    condenser, 

431. 

Pinther  powdered-coal  burner,  49. 
Pipe  anchors,  622. 
Pipe  bends,  619. 
Pipe,  brass,  608. 

cast-iron,  607. 

cast-steel,  607. 

copper,  608. 

mild  steel,  607. 

sizes  of  standard,  611. 
Pipe  flanges,  610. 

size  of  extra  heavy,  615. 

sizes-of  standard,  614. 
Pipe  hangers,  622. 
Pipe  supports,  621. 

Pipe  threads,  United  States  standard,  615. 
Pipes,  equation  of,  636. 

flow  of  steam  in,  632. 

flow  of  water  in,  650. 

size  of,  for  low-pressure  drips,  588. 

strength  of,  609. 
Piping,  by-pass  system  of  steam,  625.    • 

Commonwealth  Edison  Company,  Fisk 
Street  Station,  774-787. 

condenser,  642. 

Des  Moines    City   Railway  Company, 
631. 

duplicate  system  of,  624. 

feed-water,  648. 

Heyworth  Building,  Chicago,  626. 

loop  header  system  of,  624. 

Manhattan  Elevated,  New  York,  627. 

Paul  heating  system,  643. 

Princeton  University,  623. 

specifications  for,  760. 

steam,  622. 

Webster  heating  system,  641. 

West  Albany  Station,  New  York  Cen- 
tral Railway  Company,  788. 
Pistons,  water,  530. 
Pitot  tubes,  253. 


INDEX 


897 


Plungers,  pump,  530. 
Ponds,  cooling,  454. 
Pop  safety  valves,  664. 
Positive  injectors,  547. 
Powdered  coal,  44. 
Powdered-coal  burners  (see  Burners). 
Power  consumption   of  condenser  auxil- 
iaries, 449. 
Power  cost,  Boston  Elevated,  710. 

British  electric  light  and  power  plants, 
709. 

compound  engine  plants,  711. 

depreciation,  694. 

First    National    Bank    Building,    Chi- 
cago, 710-724. 

fixed  charges,  693. 

fuel,  700. 

insurance,  699. 

interest,  699. 

isolated  stations,  723. 

labor,  699. 

maintenance,  699. 

operating  charges,  699. 

simple  engine  plant,  711. 

street    railway   plants,    C.    C.    Moore, 
701-705. 

street  railway  plants,  R.  C.  Carpenter, 
715. 

street  railway  plants,  typical  U.S.,  709. 

taxes,  699. 

Power  measurement,  741. 
Powers  thermostat,  646. 
Preheating  feed  water,  economy  of,  484. 
Pressure  gauges,  735. 
Pressure  of  aqueous  vapor  for  different 

temperatures,  399. 
Pressure  regulation,  667. 
Producer  gas,  properties  of,  67. 
Properties  of  air,  462. 

of  fuel  oil,  52. 

of  gases,  57. 

of  lubricating  oils,  670,  676. 

of  steam,  873. 
Proximate  analysis,  31. 
Pulsometer,  572. 
Pump  governors,  541. 
Pumping  engines,  surface  condensing  for, 

436. 
Pumps,  air,  552-560. 

air  lift,  572. 

air,  jet  condensers,  552. 


Pumps,  air,  sizes  of,  553. 

air,  surface  condensers,  558. 

air,  theoretical  work,  560. 

boiler-feed,  524. 

classification  of,  522. 

centrifugal,  522,  523,  560-566. 

circulating,  572. 

direct-pressure,  528. 

duplex,  524. 

duty  of,  536. 

effect  of  piston  speed  on  economy,  473. 

fly-wheel,  532,  561. 

jet,  522,  523,  566. 

Marsh  boiler-feed,  527. 

multi-stage  centrifugal,  561. 

outside  packed  plunger,  531. 

performance  of  piston,  531. 

power,  543. 

rotary,  567. 

simplex,  537. 

size  of  boiler-feed,  539. 

tests  of  (see  Tests). 

triplex,  543. 

turbine,  561. 

volute,  561. 

water  pistons  for,  530. 
Purchasing  coal,  42. 

government  specifications  for,  769. 
Purification,  chemical,  feed  water,  476. 

mechanical,  479. 

thermal,  479. 
Purifiers,  live-steam,  507. 
Purifying  plants,  Anderson,  583. 

Scaife,  581. 

We-Fu-Go,  581. 
Pyrometers,  air-recording,  737. 

Bristol  thermo-electric,  737. 

Callendar  resistance,  738. 

Fery  radiation,  740. 

Wanner  optical,  738. 

Quality  of  steam,  827. 

Qualifications  for  a  good  lubricant,  671. 

Radial  brick  chimneys,  234. 
Radiation  and  minor  losses,  boilers,  39. 
Radiation  pyrometer,  740. 
Rankine  cycle,  268. 

Rate  of  combustion,  powdered  coal,  45. 
Rate  of  combustion,  relation  of  air  sup- 
ply on,  264. 


898 


INDEX 


Rate  of  depreciation,  694. 

Rate  of  driving,   effect  on   economy  of 

boilers,  106. 
Rateau  regenerator  accumulator,  380. 

six-stage  turbine  pump,  562. 

steam  turbine,  low-pressure,  376. 
Ratio    of    cooling    water    to    condensed 

steam,  406. 

Ratio  of  heating  to  grate  surface,  98. 
Ratio  of  expansion,  862. 
Reaction  turbines,  365. 
Receiver-reheaters,  287. 
Reciprocating  engines  vs.  steam  turbines, 

386. 

Records,  power  plant,  690. 
Reducing  valves,  396,  462. 
Reenforced  concrete  chimneys,  234. 
Regulation  of  steam  turbines,  353,  368. 
Regulators,  damper,  118. 

feed- water,  541. 

Repairs,  cost  of,  power  plants,  703. 
Report  of   United   States  Naval  Liquid 

Fuel  Board,  65. 

Restricted  feed,  lubrication,  675. 
Returns  tank,  603. 
Ring  oiler,  678. 
Ring  steam  jet,  246. 
Ringleman  smoke  chart,  844. 
Riveted  joints,  steel  chimneys,  223. 
Robb-Mumford  boiler,  73. 
Robins  belt  conveyor,  192. 
Rochester  forced-feed  lubricator,  685. 
Roney  stoker,  134. 
Rope  brake,  856. 
Rotary  pumps,  523,  567. 
Rowe  feed-water  regulator,  542. 

Safety  valve,  dead  weight,  663. 

level,  563. 

pop,  564. 
Safety  valves,  capacity  of,  565. 

rules  for  loading,  564. 
Sampling  coal,  828. 
Sarco  CO2  recorder,  744. 
Saturated-air  surface  condenser,  430. 
Saturated-steam  tables,  893. 
Sawdust  as  fuel,  18. 
Scaife  system  of   feed-water  purification, 

481. 
Scale,  analyses  of  boiler,  473. 

influence  on  heat  transmission,  472. 


Schmidt,  independently  fired  superheaters, 

167. 

Schutte  ejector  condenser,  410. 
Schwartzkopff  powdered-fuel  burner,  50. 
Scotch  marine  boiler,  72. 
Screwed  fittings,  610. 
Seaton  coal  valve,  205. 
Separating  calorimeters,  745. 
Separators,  575-586. 

Austin,  579. 

baffle-plate,  579. 

Baum,  583. 

Bundy,  579. 

centrifugal,  578. 

classification  of,  576. 

direct,  580. 

exhaust  steam,  581. 

Hoppes,  577. 

Keystone,  578. 

live  steam,  576. 

location  of,  580. 

mesh,  579. 

oil,  tests  of,  582. 

reverse  current,  577. 

Stratton,  578. 
Settings,  smokeless,  124- 

standard,  68- 
Shone  ejector,  604. 
Siegrist  oiling  system,  686. 
Simple  engines,  291. 
Simplex  coal  valve,  204. 
Single  vs.  double-acting  engines,  291. 
Siphon  condensers,  408. 
Siphon  traps,  595. 
Skimmer,  Buckeye,  118. 
Slip  expansion  point,  621. 
Smoke  observation,  844. 
Smoke  prevention,  124- 
Smoke  chart,  844. 
Smoke  recorders,  749. 
Smoke,  visible,  loss  due  to,  32. 
Smokeless  furnaces,  124- 

balanced  draft,  264. 

Burke's,  151. 

Chicago     Smoke     Inspection     Depart- 
ments, 143. 

Dutch  oven,  141. 

fire  tile,  128. 

Hein rich's,  249. 

Kent's  wing  wall,  149. 

Parsons,  248. 


INDEX 


899 


Smokeless  steam  jets,  245-250. 

stokers,  125-140. 
"  twin  fire-arch,  143. 

Wooley,  148. 
Solid  lubricants,  671. 

South  Side  Elevated  Ry.  Co.,  Chicago, 
chimney  at,  221. 

coal  crusher  and  cross  conveyor  at,  190. 
Special  furnaces,  141. 
Specific  heat  of  superheated  steam,  155- 
162. 

A.  R.  Dodge,  156. 

C.  C.  Thomas,  161. 

C.  E.  Burgoon,  157. 

Knoblauch  and  Jakob,  156,  159. 

value  of  cp  at  atmospheric  pressure, 

157. 

Specific  volume  of  superheated  steam,  160. 
Specific  gravity  of  lubricating  oils,  672. 
Specifications,  boiler,  754. 

condenser,  758. 

engine,  750. 

government,  for  purchasing  coal,  769. 

piping,  760. 

Speed,  influence   on   piston   pump   econ- 
omy, 2. 

influence  on  engine  economy,  290. 
Split  bridge  wall,  151. 
Spray  fountain,  455. 
Sprinkling  stokers,  141. 
Stability  of  brick  chimneys,  231. 
Stability  of  steel  chimneys,  223. 
Standard  method  of   starting  and   stop- 
ping boiler  tests,  825. 
Station  load  factor,  691. 
Steam  boilers,  68. 
Steam  engines,  267- 
Steam,  flow  of,  in  pipes,  632. 
Steam  jets,  245-250. 
Steam  loop,  600. 
Steam  mains,  629. 
Steam  piping,  606,  668. 
Steam,  properties  of,  878. 
Steam  pumps,  522-541. 
Steam  separators,  575-605. 
Steam,  specific  heat  of  superheated,  130- 

155. 

Steam  traps,  575-605 
Steam  turbines,  327. 
Steel  concrete  chimneys,  234. 
Steel  chimneys,  219. 


Step  bearing,  356. 
Stirling  boiler,  87. 
Stirling  superheater,  164. 
St.  John's  steam  meter,  734. 
Stokers,  125-140. 

American  underfeed,  139. 

Babcock  &  Wilcox,  128,  132. 

chain  grates,  126. 

cost  of,  151. 

Green  chain  grate,  127. 

Jones  underfeed,  138. 

Murphy,  137. 

Roney,  134. 

Wilkinson,  136. 
Stop  valves,  655. 
Storage,  coal,  181. 

oil,  64. 

powdered  coal,  45. 
Stratton  separator,  578. 
Sulphur  in  fuel,  25,  40. 
Sulphur  dioxide,  properties  of,  26. 
Sulzer   engine     for   superheated    steam, 

319. 

Superheat,  limit  of,  154. 
Superheated  steam,  153-180. 

economy  of,  steam  engine,  313. 

economy  of,  steam  turbines,  386. 

properties  of,  180. 

specific  heat  of,  155. 

specific  volume  of,  153,  160. 
Superheating  moisture  in  air,  loss  due 

to,  37. 

Superheating  surface,  extent  of,  170. 
Superheaters,  163- 

Babcock  &  Wilcox,  163. 

Foster,  165,  168. 

independently  fired,  166. 

Schmidt,  167. 

Schwoerer,  166. 

Stirling,  164. 

tests  of  independently  fired  vs.  flue 
fired,  176. 

tests,  miscellaneous,  176-179. 

yearly  expense  for  repairs,  cast-iron, 

171. 

Supports,  pipe,  621. 
Surface  blow,  117. 
Surface  condensers,  416-430. 
Surface,  cooling,  condensers,  421. 
Surface,  heating,  feed-water  heaters,  496. 
Surface,  heating,  superheaters,  170. 


900 


INDEX 


Tan  bark  as  fuel,  18. 
Tank,  blow-off,  117. 
Tank,  returns,  602. 
Taxes,  pqwer  cost,  693. 
Telescopic  oiler,  677. 
Temperature-entropy  diagram,  864,  881 . 
Temperature  of  combustion,  27. 
Temperature  measurements,  736. 
Temperature  regulators,  646. 
Terry  steam  turbine,  345. 
Tests  of  blowers,  257. 
Tests  of  boilers: 

A.S.M.E.  code,  822. 

coal  burning,  100. 

evaporation,  Armour  Glue  Works,  105. 

influence  of  draft  on  efficiency,  106. 

influence  of  rate  of  combustion  on  air 
supply,  264. 

influence  of  rate  of  driving  on  capac- 
ity, 107. 

influence  of  size  of  coal  on  capacity,  40. 

influence  of  thickness  of  fire  on  effi- 
ciency, 109. 

oil  fuel,  62. 

powdered  coal,  46. 
Tests  of  burners,  oil  fuel,  62. 
Tests  of  chimney,  100-foot  steel,  213. 
Tests  of  condenser  auxiliaries,  449. 
Tests  of  condensers: 

dry  air,  429. 

evaporative,  434. 

Fennel  saturated  air,  432. 

Weigh  ton  surface,  421. 
Tests  of  cooling  towers,  468. 
Tests  of  economizers,  515. 
Tests  of  engines,  A.S.M.E.  code,  846. 

binary  vapor,  324. 

compound     condensing    vs.     non-con- 
densing, 304. 

compound  vs.  simple  high-speed,  303. 

condensing,  increase  in  power  due  to, 
307. 

Corliss,  369. 

5500-h.p.  engine  at  Waterside  Station, 
309. 

four-valve  vs.  single-valve  high-speed, 
298. 

friction,  284. 

increasing  back  pressure,  283. 

increasing  initial  pressure,  286. 

Reeves  simple  engine,  297. 


Tests  of  engines,  simple  high-speed  engines, 
saturated  steam,  292. 

tables  of,  296,  306. 

superheated  steam,  compound,  317. 

superheated  steam,  influence  of  degree 
of  superheat,  322. 

superheated  steam,  influence  of  size  of 
engine,  332. 

superheated    steam,     record    perform- 
ance of,  321. 

superheated  steam,  triple,  318. 
Tests  of  furnace,  relation  of  gas  compo- 
sition to  temperature,  36. 
Tests  of  injectors,  549. 
Tests  of  jets,  steam,  248. 
Tests  of  oil  burners,  62. 
Tests  of  oil  separators,  572. 
Tests  of  oils,  fuel,  52. 
Tests  of  oils,  lubricating,  676. 
Tests  of  pipe  coverings,  617. 
Tests  of  pumps: 

air  lift,  572. 

boiler  feed,  534. 

centrifugal,  De  Laval,  568. 

centrifugal,  Lea-Degan,  569. 

centrifugal,  Morris,  564. 

centrifugal,  Worthington,  568. 

direct-connected  triplex,  544. 

duplex  fire  pump,  533. 

geared  triplex,  545. 

rotary,  570. 

Tests  of  separators,  576. 
Tests  of  spray  fountain,  455. 
Tests  of  superheaters,  176-181. 
Tests  of  turbines,  390. 
Thermal  efficiency  of  engines,  273. 
Thermal  purification  of  feed  water,  479. 
Thermo-electric  pyrometers,  736. 
Thermometers,  classification  of,  731. 
Thermostat,  Powers,  646. 
Thickness  of  fire,  109. 
Thickness  of  walls,  brick  chimneys,  226. 
Thomas,    specific    heat    of    superheated 

steam,  161. 

Throttling,  calorimeter,  746. 
Throttling,  moisture  evaporated  by,  312. 
Throttling  vs.  automatic  cut-off,  310. 
Tile,  "Economy,"  131. 
Tile-roof  furnaces,  63. 
Tilden  damper  regulator,  119. 
Tomlinson  condenser,  415. 


INDEX 


901 


Towers,  cooling,  456-470. 

Alberger,  459. 

Barnard,  457. 

test  of,  468. 

theory  of,  460. 

Worthington,  458. 
Traps  (steam),  588-599. 

Acme,  590. 

bowl,  591. 

bucket,  590. 

Bundy,  591. 

classification  of,  588. 

Columbia,  592. 

differential,  594. 

Dunham,  593. 

expansion,  592. 

Flinn,  595. 

float,  589. 

Geipel,  593. 

Heintz,  594. 

location  of,  596. 

McDaniel,  589. 

return,  596. 

Traveling  coal  hoppers,  203. 
Traveling  grates,  126. 
Triumph  powdered-coal  furnace,  50. 
try  cocks,  3,  120. 
Tube  cleaners,  121. 
Tupper  grate  bar,  115. 
Turbines  (steam),  327. 

advantages  of,  382. 

Allis-Chalmers,  372. 

correction  factors,  387. 

cost  of,  392. 

Curtis,  350. 

De  Laval,  331. 

Double-flow,  369. 

economy  of  space,  382. 

efficiency  of,  384. 

elementary  theory,  328. 

Hamilton-Holzworth,  362. 

impulse,  331. 

influence  of  superheat,  393. 

influence  of  vacuum,  395. 

Kerr,  3-16. 

low-pressure,  376. 

overload  capacity,  384. 

reaction,  365. 

regulation,  384. 

simplicity  of,  382. 

Terry,  345. 


Turbines  (Steam),  tests  of,  390. 

Westinghouse-Parsons,  365. 
Turner  oil  filter,  688. 
Twin  fire-arch  furnace,  142. 

Ultimate  analysis,  31. 

Underfeed  stokers,  138. 

Unit  of  evaporation,  88. 

Units,  conversion,  877. 

Universal  calorimeter,  746. 

Useful  life  of  power-plant  appliances,  694. 

Vacua,  increase  of  power  due  to  increas- 
ing, 307. 
Vacua,  influence  of,  on  economy  of  engines, 

307. 
Vacua,  influence  of  high,  on  steam  tur- 

Wnes,  395. 

Vacuum  ash  conveyor,  196. 
Vacuum  augmeriter,  Parsons,  444. 
Vacuum  chambers,  529. 
Vacuum,  degree  of,  as  affected  by  aqueous 
vapor,  405. 

drips  under,  597. 

most  economical,  451. 
Vacuum  pumps,  552. 
Vacuum  systems,  high,  441. 
Valves,  Act  on  atmospheric  relief,  666. 

Anderson  automatic  non-return,  658. 

Anderson  triple  duty  emergency,  659. 

atmospheric  relief,  665. 

back  pressure,  665. 

blow-off,  661. 

by-pass,  Westinghouse-Parsons,  369. 

check,  660. 

coal,  205. 

consolidated  pop  safety,  668. 

Crane  atmospheric,  666. 
•    Crane  hydraulic  emergency,  659. 

Davis  back  pressure,  665. 

diaphragm,  647. 

disk,  527; 

emergency,  558. 

foot,  668.   . 

Foster  back  pressure,  665. 

gate  and  globe,  655. 

Illinois  Eng.  Company's  vacuum,  643. 

Kieley  reducing,  667. 

Ludlow  angle,  gate  pattern,  657. 

non-return,  658. 

nozzle,  Curtis  turbine,  352. 


902 


INDEX 


Valves,  nozzle,  De  Laval  turbine,  332. 

nozzle,  theoretical,  334. 

Paul  vacuum,  645. 

reducing,  666. 

stop,  655. 

Webster  vacuum,  643. 
Vanes  of  Curtis  turbine,  352. 

Hamilton-Holzworth  turbine,  362. 

Westinghouse-Parsons  turbine,  365. 
Van  Stone  joint,  613. 
Vegetable  oils,  669. 
Velocity  of  steam  through  nozzles,  328. 

through  pipes,  633. 
Velocity  of  water  through  nozzles,  650. 

through  pipes,  650. 
Venturi  meter, 

Vertical  blow-off  connections,  116. 
Vertical  tubular  boilers,  69.  * 

Volatile  matter  in  coal,  23. 
Visible  smoke,  loss  due  to,  38. 
Viscosity  of  oils,  673. 
Volute  centrifugal  pump,  561. 
Volume  of  jet  condenser  chamber,  408. 

Wainwright  feed-water  heater,  493. 

Wall  brackets,  piping,  622. 

Wanner  optical  pyrometer,  738. 

Warren  fuel  oil  burner,  59. 

Washed  coals,  40. 

Water  and  boiler  scale,  analyses  of,  407. 

Water  columns,  119. 

Water-cooling  systems,  453. 

Water,  flow  of,  in  pipes,  650. 

Water,  friction  coefficient  in  clean  iron 
pipes,  652. 

Water,  height  of  lift  by  suction,  538. 

Water  pistons,  530. 

Water-softening  plants,  480. 

Water  temperatures,  feed-water  heaters, 
490. 

Water,  weight  of,  for  condensing,  406. 

Waterworks,  centrifugal  pump  charac- 
teristics for,  565. 

Weber  concrete-steel  chimneys,  235. 


Webster  feed-water  heaters,  488. 

Webster  heating  system,  406. 

Webster  vacuum  valve,  643. 

We-Fu-Go  purifying  system,  481. 

Weighing  fuel,  852. 

Weighing  water,  850. 

Weight   of   air  as   indicated   by  flue-gas 

analysis,  30. 
Weight  of, 
Weight  of    boiler  compound   necessary, 

478. 

Weight  of  guyed  steel  chimneys,  219. 
Weight  of  water  evaporated  per  square 

foot  of  heating  surface,  95. 
Weigh  ton  multi-flow  surface  condenser, 

419. 

Weiss  barometric  condenser,  411. 
West  Albany  power  station ,  N.  Y.  C.  R.  R., 

788-797. 

Western  Electric  Co.'s  Power  plant,  272. 
Westinghouse-Leblanc  condenser,  445. 
Westinghouse-Parsons    steam    turbines, 

365. 

Wet  air  pumps,  553. 
Wheeler    admiralty    surface    condenser, 

417. 

Wheeler,  C.  H.,  multi-flow  surface  con- 
denser, 443. 

White-Star  oil  filter,  688. 
Wickes,  boiler,  84. 
Wilkinson  stoker,  136. 
Williams  oil  burner,  58. 
Wire  drawing,  254. 
Wood  as  fuel,  18. 
Wooley  smokeless  furnace,  149. 
Worthington  barometric  condenser,  414. 
conoidal  pump,  test  of,  568. 
cooling  tower,  458. 
jet  condenser,  402. 
Wrought-iron  pipe,  611. 

Yonkers  power  house,  N.Y.C.R.,  piping 
for, 

Zinc,  use  of,  in  boilers,  477. 


mm 


•   ••••'  m 

m 

$m&a$ 


mm 

m  m 


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